The 


Una-Flow  Steam-Engine 


By 

Prof.  Dr.-Ing.  h.  c.  J.  Stumpf 

Technische  Hochschule,  Berlin. 


Translated  by  the 

Stumpf  Una-Flow  Engine  Company,  Inc., 

401  S.  A.  &  K.  Building. 

Syracuse  N.Y. 


Second  Edition 


1922 


SI 


Copyright   1922   by  Prof.  J.  Stumpf,  Berlin 


THE  BA  TTLE  OF  THE  ELEMENTS 

By  J.  A.  STUM  PP. 


(Chaos  1831,  Nr. 


saw  what  he  had  made  and  found  it  good, 
wrote  a  man  of  noblest  mind  and  mood. 
No  longer  with  this  doctrine  is  the  world  content, 
The  doubter  does  in  bitter  words  lament: 
One  need  but  cast  a  fleeting  glance  at  life, 
What  sees  one  there?    True  happiness?   No,  strife, 
Death,  need  and  misery  far  and  wide, 
The  elements  in  constant  war  abide, 
And  storms  of  passion  breeding  endless  hate 
Rob  life  of  peace,  till  many  curse  at  fate. 

We  ask  ourselves  why  we  are  so  surrounded 

By  raw  materials  and  forces  of  all  kinds, 

On  what  the  longing  in  our  breasts  is  founded, 

United  with  the  curious  impulse  of  our  minds? 

As  master  of  the  earth,  man  shall  create! 

All  labor  does  the  builder's  hand  await. 

To  carry  out  His  plan,  God  made  him  wise, 

That  he  might  force  and  matter  utilize. 

Such  is  the  man  whose  work  success  has  crowned, 

Who  truth  and  light  by  his  research  has  found, 

Who  tested  fire's  flame  and  water's  might, 

And  thus  their  deepest  secrets  brought  to  light: 

Who  through  their  elemental  strife  conceived, 

Instead  of  ruin,  mankind's  gain  achieved. 

Nature's  forces  he  has  sought, 

And  under  his  control  has  brought. 

The  foes  who  storm  with  rage  and  hate, 

He  keeps  by  thin  walls  separate. 

Around  the  boiler  roars  the  flame, 

The  seething  waves  within  to  tame, 

Who,  in  revenge,  their  enemy  to  reach, 

Strive  through  the  prison's  walls  to  force  a  breach. 

A  polished  rod  ascends,  by  magic  trained, 
Propelled  by  steam  within  a  pipe  contained. 
But  lo!  into  the  angry  steam  so  bold, 
Now  pours  a  rage-appeasing  flood  of  cold; 
Down  slides  the  rod,  but  in  an  instant  back 
Pursued  again  by  live  steam  in  its  track. 
The  shining  steel  glides  to  and  fro 
And,  driving  other  parts,  all  show 
A  striving  to  one  goal.    The  great  machine 
Obeys  the  master's  mind,  it  may  be  seen. 

How  many  nature's  wondrous  course  deride, 
And  what  they  do  not  grasp,  they  claim  unproved, 
The  man  of  science  does  regard  with  pride 
How  parts  and  whole  in  best  accord  are  moved. 


520223 


ERRATA 


Page 
70-71 
124 
125 
125 
206 
234 

234-235 
303 


Line  or  Fig. 
Fig.  1 
Line  17 
Heading 
Curve 
Last  line 
Line  11 
Fig.  13 
Title 


Change  from 
Max.  Continuous  L.H.P. 
Omit— "&  T.  Hall" 
Omit— "&  Hall" 
Corliss 
Lead 
Fig.  14 

German  State  Rys. 
Omit  "Triple  speed" 


To 
Max.  Cont.  I.H.P. 


Counterflow 

Load 

Fig.  13 

Russian  State  Rys. 


TABLE    OF    CONVERSION    FACTORS 


1  mm. 

1  cm. 

1  m. 

1  km. 

1  sq.  cm. 

1  sq.  m. 

1  kg. 

I  metric  ton 

1  metric    ton-km. 

1  m.  kg. 

1  kg.  per  sq.  cm. 

1  atmosphere 

1  metric  H.P. 

1°  C. 

Temp,  in  °  F. 

1  calorie 

1  cal.  per  kg. 

1  cal.  per  I.H.P.  hr.  (metric) 

1  kg.  per  I.H.P.  hr.  (metric) 


0.039  in. 
0.394  in. 
3.28  ft. 
1.609  miles 
0.155  sq.  in. 
10.76  sq.  ft. 
2.205  Ibs. 
2204.6  Ibs. 
0.685  ton-mile 
7.233  ft.  Ibs. 

14.22  Ibs.  per  sq.  in. 

0.9863  H.P. 

1.8°  F. 

1.8  X  temp,  in  °C.  +  32 

3.968  B.T.U. 

1.80  B.T.U.  per  Ib. 

4.024  B.T.U.  per  I.H.P.  hr. 

2.236  Ibs.  per  I.H.P.  hr. 


GENERAL    INDEX 


Acceleration,  valve,  88,  308 
Admission,  14  to  16 
A.  E.  Cr.,  69 

Area,  inlet  valve,  50  to  58 
Area,  exhaust  port,  60,  69 
Auniliary  exhaust  valves,  46,  47,  167,  177, 
208,  225,  264 

B 

Balancing,  72,  297 
Bearings,  proportions  of,  73 
Belt,  exhaust,  12,  69 
Blast,  112,  248 
Bleeding,  187 

automatic  control  for,  190 
Bonnet,  valve,  Lentz,  87 

Stumpf,  128 


Cage,  valve,  2,  80,  170 
Cam 

oscillating,  89,  143,  144,  164,  167,  170, 
172,  206,  246,  264,  270 

reciprocating,  128,  160,  208,  217,  231, 
257,  277,  279,  287,  291,  292 

revolving  tapered,  208,  218,  297,  300 

revolving  stepped,  271,  293,  294 
Compounding,  2,  5,  190,  194 
Compression,  best  length  of,  17  to  34 
Compressor,  218 

combined  steam  and  air  cylinder,  213 
Condensation,  initial,  1 
Condenser,  68 

jet,  69 

surface,  276,  294 

Westinghouse-Leblanc,  69 
Condensing  engine,  68,  126 
Connecting  rod,  173,  264 
Consumption,  steam, 

lowest   for  various   compressions,   17 
to  26 

lowest  for  various  clearances,  27  to 

29 

Convection,  losses  due  to,  105 
Corliss  una-flow  engine,   182 
Crank,  74,  75 

center,  71,  132 

Crank  pin,  proportions  of,  73 
Crosshead,  171,  294 

pin,  proportions  of,  73 
Cut-off,  range  of,  126 
Cylinder,  boring  of,  93,  155 

material  of,  92 

D  • 

Diagram,  indicator,  166,  175,  203,  269,  286 

Diffusor,  109,  110,  249 

Doerfel,  172 

Draft,  smoke  box,  112,  113,  248 

Drop,  pressure,.  51 

Dual  clearance,  177 

E 

Eccentric  gear,  single,  150 
Efficiency,  mechanical,  71 


Engines 

Ames  Iron  Works,  160,  161 
automotive,  272 
blowing,  213,  223 
Borsig,  A.,  245 

Burmeister  &  Wain,  132,  137,  278 
compressor,  213,  218 
Corliss  una-flow,  182 
Dehne,  A.  L.  G.,  218 
Ehrhardt    &    Sehmer,    137,    139,    143, 
195,  196,  202 

Elsaessische  Maschinenfabrik,  129 
Erste  Bruenner   Maschinenfabrik, 
129,  262 

Filer  &  Stowell  Co.,  178,  181 
Frerichs  &  Co.,  J.,  273 
Goerlitzer  Maschinenbauanstalt,  144 
Gutehoffnungshuette,  209 
Harrisburg  Fdy.  &  Mch.  Works,  177, 
178 

hoisting,  207 

Hungarian  State  Rys.,  264 
Kingsford  Fdy.  &  Mch.  Works,  292 
Kolomna  Engine  Works,  225,  234,  237, 
238,  262 

Linke  Hoffmann  Works,  218 
List,  Gustav,  213 
Locomotive,  225 
Marine,  272 

Maschinenbauanstalt  Breslau,  232 
Maschinenfabrik  Augsburg-Nuern- 
berg   (M.  A.  N.),   144 
Maschinenfabrik  Badenia,  252,  255 
Maschinenfabrik  Esslingen,  155 
Mesta  Machine  Company,  206,  223 
Musgrave  &  Sons,  Ltd.,  143 
Neuruppin-Kremmen-Wittstock  Ry., 

240 

Nordberg  Mfg.  Co.,  172,  177 
Northeastern  Ry.  of  England,  240 
Northern  Ry.  of  France,  234 
portable,  252 
pumping,  213,  218 
Robey  &  Co.,  262 
Rolling  mill,  139,  195 
Schmid,  Karl,  294 

Schweizerische  Lokomotivfabrik,    232 
Skinner  Engine  Co.,  168  to  171 
Soumy  Machine  Works,  159 
stationary,  127 
Stork  &  Co.,  144 

Sulzer  Bros.,   10,  76,  150,  151,  155 
Vulkan  Engine  Works,  225,  227,  240, 

274 
Worthington   Pump   &   Mchy.   Corp., 

218 

Ejector  effect,   110,   112   to   117,  245,  249, 
250,  269 

saving  due  to,  117 
Expansion,  cylinder,  128 
Experiments,  Prof.  Naegel's,  118  to  124 


Friction 

of  driving  parts,  77 
H.P.,  75-77 
loss  due  to,  71 


single  eccentric,  150 

valve,  cam  (locomotive),  230,  245 

double-speed,  295,  303 

Gooch,  207 

Klug,  275,  278,  280,  290 

link,  disadvantage  of,  16 

Marshall,  237 

Saeuberlich,  274 

Skinner  auxiliary  exhaust,  171 

Stumpf,  128,  129 

Walsnhaert,  22,  230,  247,  285 

Zvonirek,  143,  195 
Governor,  flywheel,  252 
Gueldner,  110 

I 

Inertia,  curves  of,  72 
Insulation,  105 

J 

Jacketing,  27 

effect  of,  2,  3,  6,  10,  11,  12,  33 

Sulzer's  tests  on,  11  • 
Jackets,  proportions  of,  12 
Joints,  piston  ring,  94,  95 


Lagging,   105 

Lap,   128,  202,  207,  240 

Lead,  exhaust,  5,  61,  108,  239,  269 

Leakage 

losses  due  to,  79 
valve,  effect  of,   85 

Lentz  packing,  100 

Liner,  160 

Locomobile,  252 

Locomotive,  225 
Borsig,  245 
Kolomna  Engine  Works,  225,  234,  237, 

238 

Maschinenbauanstalt  Breslau,  232 
Neuruppin-Kremmen-Wittstock  Ry., 

240 

Northeastern  Ry.  of  England,  240 
Northern 'Ry.  of  France,  234 
Schweizerische  Lokomotivfabrik,  232 
Vulkan  Engine  Works,  225,  227,  240 
pistons  for,  99 
,  three  cylinder,  116,  240 

Losses 

friction,  71 

incomplete  expansion,  108 

leakage,  '.^ 

radiation  and  convection,  105 

surface,  1,  4 

throttling,  49 

volume,  13 


Lubrication 

cylinder,  99,  155,  227,  248,  273 
driving  parts,  78 
valve  gear,  87,  128 

M 

Machining 

of  clearance  surfaces,  2 
of  cylinders,  93,  155 
of  pistons,  93 

Marine  engines 

Burmeister  &  Wain,  278 

Frerichs,  &  Co,  J,  273 

Kingsford  Fdy.  &  Mch.  Works,  292 

Schmid,  Karl,  294 

Vulkan  Engine  Works,  274 

N 

Nagel's  experiments,  118  to  124 
Nozzle,  109,  116,  249,  269 


Parts 

driving,  proportions  of,  71,  73,  75,  76 
reciprocating,  71,  72 

Packing,  piston  rod,  100  to  104 

Pin 

crank,  proportions  of,  73 
crosshead,  proportions  of,  73 

Pipe 

blast,  111,  112,  115,  248,  249 
exhaust,  109,  110 
steam,  105 

Piston 

cast  steel,  78,  227 
expansion  of,  93 
floating,  77,  92  to  94,  99,  155 
machining  of,  93,  155,  227 
overrunning  of,   94 
radial  clearance  of,  93,  155 
self-supporting,  77,  92,  155 
shoes,  77,  92,  155,  240 
three-piece,  227 
two-piece,  77,  240,  264 

Portable  engines 

Erste  Bruenner  Maschinenfabrik,  262 
Hungarian  State  Rys.,  264 
Kolomna  Engine  Works,  262 
Maschinenfabrik  Badenia,  252,  255 

Ports,  exhaust,  5,  69 
area  of,  60,  113 

Pressure,  back,  29  to  33,  42   ' 

critical,  34  to  43 
Pump,  tube,  223 
Pumping  engines 

List,  Gustav,  213 

Worthington  Pump  &  Mchy.  Corp., 
218 


R 

Radiation,  losses  due  to,  105 

Rating,  load,  71 

Relief,  compression,  for  starting,  203,  207, 

229,  230,  240,  247,  275 
Resilient  valve,  calculation  of,  81 
Rings 

piston,  94 

hammered,  96 

overrunning  of,  97,  98 

proportions  of,  96 

wear  of,  98 
Rod,  connecting,  72,  173,  263 

tail,  77,  93,  206 
Rolling  mill   engines 

Ehrhardt  &  Sehmer,  195,  196,  202 

Mesta  Machine  Co.,  206 

Schlick,  297 

Seats,  valve,  80,  85,  86 

Separators,  oil,  69 

Series  arrangement  of  parts,  85 

Shaft,   crank,  proportions  of,  73 

Shoes,  piston,  77,  92,  155,  240 

Slap,  piston,  94 

Speed,  piston,  71 

Stack,  smoke,  112 

Stationary  engines 

Ames  Iron  Works,  160,  161 
Burmeister  &  Wain,  132,  137         . 
Ehrhardt  &  Sehmer,  137,  143 
Elsassische  Mascheninfabrik,  129 
Erste  Brunner  Maschinenfabrik,  129 
Filer  &  Stowell  Co.,  178,  181 
Gorlitzer  Maschinenbauanstalt,  144 
Harrisburg  Fdy.  &  Mch.  Works,  177, 

178 
Maschinenfabrik  Augsburg- Nurn berg, 

144 

Maschinenfabrik  Esslingen,  155 
Musgrave  &  Sons,  Ltd.,  143 
Nordberg  Mfg.  Co.,  172,  177 
Skinner  Engine  Co.,  168  to  171 
Soumy  Machine  Works,  159 
Stork  &  Co.,  144 
Sulzer  Bros.,  10,  76,  150,  151,  155 


Stepped  cams,  271,  293,  294 

Stodola,  110 

Strahl,  112 

Superheater,  258 

Superheating,  effect  of,  2,  3,  5,  8,  273 

Surface  clearance,  1  to  3,  7 

Surface  clearance,  minimum,  1 

T 

Temperature 

steam,  experiments  on,  118  to  124 

high,  8 
Tests,  engine,  129,  137,  144,  157,  161,  234, 

236,  265 

Throttling,  losses  due  to,  49 
Tube  pump,  203 

V 

Vacuum,  high,  68 
Valves 

clearance,  45,  46 
Corliss,  79,  183  to  186 
double  beat,  leakage,  of,  79 
locomotive,  230 
resilient,   calculation   of,   81 
tightness  of,  79,  80 
two-piece,   168 

exhaust,    auxiliary,    46,    47,    166,    167, 
177,  208,  225,  264 

automatic,  166,  264 
inlet,  area  of,  50  to  58 

leakage  of,  85 
piston,  79,  177,  178,  196,  210,  239,  240, 

275 
single  beat,  86,  245,  247,  262,  272,  295, 

303 
slide,  79 

locomotive,  237 

Valve  spring,  calculation  of,  88,  308 
Volume 

clearance,  additional,  44  to  46,  126 
clearance,  %  of,  48,  126,  127,  228 

w 

Walschaert  gear,  22,  230,  247,  28.5 
Westinghouse-Leblanc,  69 

z 

Zeuner,  111 

Zvonicek  gear,  143,  195 


Preface. 

The  second  edition  of  this  book  represents  a  complete  revision  of  the  first 
one,  little  of  which  remains.  The  first  edition  contained  a  good  many  opinions  in 
addition  to  facts  and  was  intended  rather  for  the  purpose  of  defending  the  una- 
flow  engine  against  antiquated  theories  and  attacks.  In  the  meantime  the  una- 


Fie.  1. 


flow  principle  has  been  widely  tried  out  and  scientifically  investigated.  It  has 
become  an  accomplished  fact  and  is  in  common  use.  This  book  therefore  contains 
scientific  proof  from  an  objective  point  of  view,  as  well  as  a  description  of  the 
development  of  the  una-flow  engine. 


In  the  opening  chapters  of  the  book  the  different  losses  of  the  steam  engine 
are  investigated.  The  causes  and  effects  are  defined,  as  well  as  the  relations  be- 
tween them,  and  the  manner  is  pointed  out  in  which  the  minimum  value  of  each 
loss  is  obtainable.  After  considering  all  the  seven  different  losses  occurring  in 
a  steam  engine,  the  question  is  asked  as  to  how  a  steam  engine  must  be  designed 
in  order  to  have  a  minimum  total  of  all  the  seven  losses.  In  answer  to  this  query 
two  different  designs  are  presented,  one  being  a  stationary  una-flow  engine  with 
single-beat  valves  for  condensing  operation,  and  the  other  a  una-flow  locomotive 
also  equipped  with  single-beat  valves.  On  the  former,  the  use  of  single-beat  valves 
was  made  possible  by  the  use  of  a  double  speed  valve  gear  (see  Chapter  VI). 

Both  types  of  engines  were  developed  during  the  year  1920.  Lower  steam 
consumption  figures  than  those  given  by  the  best  multi-stage  engines  are 


Fig.  2. 

obtainable  with  these  types  for  both  saturated  and  superheated  steam.  The  expe- 
rience gained  with  una-flow  engines  in  widely  different  fields  and  under  the  most 
varying  conditions  was  utilized  in  the  design  of  these  engines  to  the  fullest  pos- 
sible extent. 

A  number  of  Chapters  is  devoted  to  the  description  of  this  development  in 
all  its  phases. 

The  novelty  in  this  case  is  the  single-beat  valve,  which  has  so  far  been  used 
only  in  internal  combustion  engine  practice.  All  previous  attempts  to  apply  it  to 
steam  engines  have  miscarried.  The  application  of  this  type  of  valve  to  the  una- 
flow  engine  represents  figuratively  the  keystone  in  the  development  of  the  latter. 
The  fundamental  conformity  between  the  new  una-flow  engine  and  the  two-stroke 
internal  combustion  engine  is  surprising.  This  refers  to  the  uni-directinal  flow,  the 
single  stage  expansion,  the  piston-controlled  exhaust  and  the  single-beat  inlet  valve. 
?!  Surprising  also  is  the  close  agreement  in  the  essential  parts  of  the  cylinder 
between  the  latest  design  shown  in  Fig.  3  (see  Chapter  VI)  and  the  first  original 


sketch  of  a  una-flow  engine  made  in  the  year  1900  and  reproduced  in  Fig.  2,  which 
also  incorporates  single-beat  valves. 

Since,  as  Descartes  says,  doubt  may  be  considered  the  origin  of  every  philo- 
sophy, the  question  regarding  the  doubt  which  originated  the  una-flow  philosophy 
may  well  be  asked.  This  doubt  arose  in  the  year  1896  during  the  starting  up  of 
two  pumping  engines  designed  by  myself  for  the  Pope  Mfg.  Co.,  of  Hartford,  Conn. 
(Fig.  1).  These  were  vertical  triple  expansion  engines  with  Corliss  valves  and  a 
central  condensing  system,  in  which  everything  then  considered  good  practice 
was  carried  to  the  extreme  limit.  This  resulted  in  a  very  complicated  construction 
which  appeared  to  me  to  be  a  sign  of  weakness.  The  doubts  which  then  arose 
in  my  mind  eventually  led  to  the  sketch  shown  in  Fig.  2  during  the  year  1900. 
The  construction  of  steam  turbines  of  several  stages,  which  began  at  that  time, 


Fig.  3. 

was  developed  along  the  lines  of  pure  uni-directional  flow,  arid  this  brought  up 
the  question  whether  it  would  not  be  possible  to  raise  the  reciprocating  steam 
engine  to  the  same  thermal  plane  as  the  turbine  by  the  use  of  the  una-flow  prin- 
ciple. The  application  of  the  una-flow  action  of  the  turbine  to  the  steam  engine, 
although  in  a  somewhat  imperfect  manner,  by  properly  designing  the  cylinder, 
valve  gear,  steam  jackets  and  condenser  connection,  etc.,  finally  led  to  the  una- 
flow  engine  with  single-beat  valves  as  shown  in  Fig.  3.  The  object  in  view  was 
the  attainment  of  the  minimum  total  of  all  the  seven  different  losses  of  the  steam 
engine,  as  well  as  the  utmost  simplicity  and  reliability  of  operation.  This  goal 
now  seems  to  have  been  fully  reached.  The  fact  that  the  una-flow  engine  pos- 
sesses the  uni-directional  flow  in  common  with  the  steam  turbine  and  has  a  con- 
structional basis  similar  to  that  of  the  two-stroke  internal  combustion  engine, 
may  be  cited  in  support  of  this. 

The  design  and  adaptation  of  the  una-flow  engine  to  different  requirements 
and  conditions  of  service  represents  an  immense  amount  of  work,  in  which  I  received 


the  full  support  of  my  assistents  as  well  as  that  of  Mr.  Rosier  of  Miihlhausen, 
Alsace,  Mr.  Arendt  of  Saarbriicken,  Prof.  Bonin  of  Aachen,  Mr.  Dutta  of  London, 
and  Drs.  Mrongovius  and  Meineke  of  Berlin. 

To  the  splendid  support  of  the  last  four  gentlemen  may  be  attributed  the 
positive  developments  of  the  chapters  on  volume  loss,  throttling  loss,  exhaust 
ejector  action,  the  una-flow  locomotive,  and  the  valve  gear  with  double  speed 
lay  shaft. 

I  am  particularly  indebted  to  those  gentlemen  who  took  up  my  proposals  at 
a  time  when  no  one  would  yet  believe  in  the  una-flow  engine,  namely,  Prof.  Nol- 
tein  of  the  Technical  Hochschule  at  Riga,  Messrs.  Hnevkovsky  and  Smetana 
of  Brtinn,  Mr.  Lamey  of  Miihlhausen,  Alsace,  Mr.  Mtiller  of  Berlin,  and  Mr.  Schiller 
of  Grevenbroich. 

Berlin,   January  1921. 

J.  Stumpf. 


Index. 

Page 

Preface V 

I.  Steam  Engine  Losses 1 

1  a.  Losses  due  to  Cylinder  Condensation 1 

1  b.  The  Una-Flow  Arrangement  as  a  Means  for  reducing  Surface  Losses. 

Jacketing  of  the  Cylinder 4 

2  a.  The  Influence  of  the  Clearance  Volume  upon  the  theoretical  Steam  Con- 

sumption (Volume  Loss).    Critical  back  Pressure 13 

2b.  Additional  Clearance  Space 44 

3  a.  Losses  due  to  Throttling.  Determination  of  Inlet  and  Exhaust  Port  Areas  49 
3b.  Relation  between  the  Una-Flow  Engine  and  the  Condenser     *   .   .   .   .  68 

4.  Lossesdueto  Friction  (mechanical  Efficiency).  Dimensioning  of  drivingParts  71 

5.  Losses  due  to  Leakage.    Valves,  Pistons,  Piston  Rod  Packings  ....  79 

6.  Losses  due  to  Radiation  and  Convection 105 

7.  Losses  due  to  incomplete  Expansion.     The  Exhaust  Ejector  Effect  .    .  108 

8.  Prof.  Nagel's  Experiments 118 

II.  1.     The  Una-Flow  Stationary  Engine 126 

2.  The  Una-Flow  Corliss  Engine 182 

3.  .  The  Una-Flow  Engine  arranged  for  Bleeding 187 

4.  The  Una-Flow  Rolling  Mill  Engine 195 

5.  The  Un^-Flow  Hoisting  Engine 207 

6.  Una-Flow  Engines  for  driving  Air  Compressors,  Pumps,  etc 213 

III.  The  Una-Flow  Locomotive 226 

IV.  The  Una-Flow  Locomobile  and  Portable  Engine 252 

V.  The  Una-Flow  Marine  Engine 273 

VI.  The  Una-Flow  Engine  with  single-beat  Valves  and  double-speed  Lay  Shaft  303 

Summary 316 


I.  Steam  Engine  Losses. 


The  losses  in  a  steam  engine  may  b.e  classified   as   follows: 

1.  Losses   due  to   cylinder   condensation   (surface  loss). 

2.  Losses  due  to  the  volume  of  the  clearance  space  (clearance  volume  loss). 

3.  Loss   due   to   throttling   or  wire  drawing. 

4.  Friction   loss. 

5.  Loss   due  to  leakage. 

6.  Loss   due  to   heat  radiation  and   convection. 

7.  Loss   due  to  incomplete  expansion. 

> 

la.  Losses  due  to  Cylinder  Condensation 

(Surface  Loss). 

The  amount  of  initial  (or  cylinder)  condensation  (termed  surface  loss  in  the 
following)  is  determined  by  the  size,  kind  and  arrangement  of  the  harmful  sur- 
faces, by  the  steam  jacket,  by  the  quality  of  steam  passing  these  surfaces,  by  the 
temperature  gradient  and  the  number  of  stages  used,  by  the  amount  and  period 
of  the  steam  flow  and  the  path  of  the  steam  through  the  cylinder  (counterflow 
or  una-flow).  Initial  condensation,  which  is  usually  over  by  the  end  of  admission, 
is  caused  by  the  clearance  surfaces  and  increased  by  any  moisture  carried  over 
with  the  steam.  The  ability  of  these  surfaces  to  receive  and  give  off  heat  forms 
a  kind  of  heat  bypass,  with  a  corresponding  loss  to  the  cycle.  Part  of  the  steam 
condenses  during  admission  and  re-evaporates  during  exhaust  and  the  last  part 
of  expansion. 

The  harmful  surfaces  comprise  the  inner  surfaces  of  the  cylinder  and*  the 
inwardly  exposed  surfaces  of  piston,  piston  rod  and  steam  distributing  parts. 
Surfaces  which  are  continually  exposed  even  in  the  dead  center  position  of  the 
piston  may  be  termed  harmful  surfaces  of  the  first  order,  and  those  which  are 
progressively  uncovered  by  the  piston  during  its  motion,  harmful  surfaces  of  the 
second  order. 

The  former  usually  cause  the  essentially  greater  part  of  the  surface  loss.  Since 
the  amount  of  surface  loss  is  determined  by  the  extent  of  the  harmful  surfaces, 
the  latter  should  be  kept  as  small  as  possible  and  should  also  be  machined.  A 
good  many  designers  pay  attention  only  to  the  amount  of  clearance  volume  without 
considering  its  surface.  The  minimum  harmful  surface  of  the  first  order  comprises 
an  area  equal  to  twice  the  cylinder  cross-section  (cylinder  head  and  piston),  and 
it  is  convenient  to  express  the  additional  surface  of  the  first  order  in  percent  of 
this  minimum  surface.  In  actual  engines  these  additional  surfaces,  which  are 
mostly  not  machined,  are  found  to  be  from  150  to  200%  of  this  minimum  harmful 

Slump/,  The  una-Flovv  steam  engine. 


surface,  although  it  is  possible  by  careful  design  to  reduce  tiiis  figure  to  3  or  5%. 
Piston  valves  with  snap  rings  working  in  separate  bushings,  as  well  as  slide  valves 
with  long  curved  ports,  which  latter  are  in  most  cases  left  rough  and  serve  for  both 
steam  admission  and  exhaust,  have  large  surfaces  which  are  especially  harmful 
on  account  of  their  very  nature  and  arrangement. 

Engines  with  separate  admission  and  exhaust  ports  are  far  better  in  this 
respect,  because  the  latter  are  usually  very  short  and  the  hot  admission  and  cold 
exhaust  steam  enter  and  leave  the  cylinder  through  separate  passages,  thereby 
avoiding  the  alternate  heating  and  cooling  of  these  surfaces  and  the  corresponding 
surface  loss  which  takes  place  in  engines  having  common  inlet  and  exhaust  ports. 

Slide  and  piston  valve  engines  with  their  perpetual  reversal  of  flow  are  sub- 
ject to  extensive  turbulence  and  heat  exchanges,  although  careful  design  can 
generally  improve  conditions.  The  Corliss  engine  may  be  considered  an  improve- 
ment by  reason  of  .the  smallness  and  different  arrangement  of  its  additional  sur- 
faces; and,  with  the  valves  in  the  heads,  the  ports  are  straight  and  short  with 
inlet  and  exhaust  separated.  Conditions  can  be  improved  still  further  if  the  exhaust 
valve,  which  forms  the  major  part  of  the  additional  surfaces,  is  made  to  fill  its 
bore  completely  and  has  a  straight  port  through  its  center  only. 

Poppet  valve  engines  in  general  have  rather  large  additional  surfaces,  espe- 
cially where  valve  cages  are  used ;  and  notwithstanding  separate  inlet  and  exhaust 
ports  the  frequent  flow  reversal  has  a  deleterious  effect.  Valve  cages  considerably 
increase  the  additional  surfaces.  Machining  of  the  latter  may  be  provided  for 
in  many  cases  by  clever  design,  thereby  reducing  their  extent  and  the  corre- 
sponding turbulence  and  surface  loss. 

Further  means  of  reducing  the  surface  loss  are:' 

1.  Jacketing, 

2.  Compounding, 

3.  Superheating, 

4.  The  una-flow  system. 

Jacketing  of  the  harmful  surfaces  is  a  further  step  in  reducing  surface  losses. 
The  heating  medium  is  usually  steam,  seldom  flue  gases.  Engines  with  great  sur- 
face losses  will  be  largely  benefited  by  jacketing,  and  single  cylinder  condensing 
engines  working  with  saturated  steam  will  show  the  greatest  gain  since  they  offer 
the  largest  scope  for  improvement.  The  effect  of  the  jacket  is  diminished  if  the 
expansion  takes  place  in  two,  three,  or  four  stages,  and  if  the  steam  is  superheated 
in  addition.  These  means  improve  the  thermal  condition  to  such  an  extent  that 
there  is  little  to  be  gained  from  jacketing.  This  applies  to  superheating,  espe- 
cially if  the  whole  working  cycle  takes  place  in  the  range  of  superheat.  Locomo- 
tives, in  which  the  steam,  when  leaving  the  cylinder  is  still  superheated,  will 
derive  no  benefit  from  jacketing.  Superheating  is  such  a  far  reaching  remedy  that 
the  number  of  stages  in  counterflow  engines  working  with  superheated  steam  has 
been  generally  reduced  from  three  to  two  for  condensing  operation,  and  from 
two  to  one  when  operating  non-condensing. 

Increased  speed,  later  cut-offs,  and  larger  units  tend  to  reduce  the  surface 
losses  and  hence  the  effects  of  jacketing.  High  speed  damps  the  temperature 
fluctuations  of  the  walls;  the  late  cut-off  raises  the  mean  wall  temperature;  and 


the  larger  size  gives  a  more  favorable  ratio  of  volume  to  surface.  For  these  reasons 
a  large,  fast  running,  and  heavily  loaded  engine  will  show  the  least  gain  from 
jacketing,  especially  if  working  with  superheated  steam  or  a  small  temperature 
gradient.  All  this  applies  to  superheated  steam  locomotives  as  an  example. 

Jacketing  has  the  greatest  effect  in  low  pressure  cylinders,  since  surface 
losses,  temperature  gradient,  and  jacket  surfaces  are  large  and  the  weight  ratio 
of  jacket  to  working  steam  is  the  most  favorable.  The  gain  from  jacketing  is 
accordingly  smaller  in  intermediate  and  high  pressure  cylinders,  and  in  many 
cases  there  is  hardly  any  in  the  latter.  Similar  conditions  prevail  in  two  cylinder 
compound  engines.  Head  jacketing  is  usually  more  effective  than  cylinder  jacketing 
because  the  cylinder  surfaces  are  temporarily  covered  by  the  piston,  and  the  oil 
film  acts  as  a  heat  insulator.  These  surfaces  of  the  second  order  cause  small  sur- 
face losses  and  consequently  show  a  smaller  gain  from  jacketing. 

Saturated  steam  is  very  bad  in  this  respect  because  the  water  particles  act 
as  heat  conductors  and  increase  the  surface  losses.  Dry  steam  is  better,  and  best 
of  all  is  superheated  steam.  Saturated  steam  is  an  excellent,  and  superheated 
steam  a  very  poor  heat  conductor.  The  action  of  the  cylinder  becomes  the  more 
adiabatic  the  more  the  superheated  region  extends  through  the  cycle.  Superheat, 
furthermore,  by  increasing  the  specific  volume,  makes  the  steam  lighter  and 
reduces  both  the  weight  per  cycle  and  the  surface  loss.  The  commonly  prevailing 
amount  of  superheat  allows  non-condensing  engines  to  work  in  the  superheated 
region  throughout  the  cycle;  but  this  is  not  the  case  with  condensing  engines, 
assuming  proper  ratios  of  expansion  in  both  cases.  In  condensing  engines  the 
low  pressure  part  of  the  cycle  always  extends  beyond  the  saturation  point. 

Superheating  is  of  such  far  reaching  effect  that  the  reduction  of  harmful 
surfaces,  their  arrangement,  and  in  many  cases  even  jacketing,  lose  their  impor- 
tance. Generally  speaking,  among  the  different  ways  of  reducing  surface  losses, 
one  or  the  other  may  be  so  effective  that  there  is  nothing  left  for  the  remaining 
ones. 

The  amount,  extent  and  kind  of  steam  flow,  especially  of  the  exhaust,  may 
have  considerable  influence  in  engines  working  with  saturated  steam.  Wet  exhaust 
steam  flowing  with  high  velocity  through  long  unfinished  ports  having  large  sur- 
faces may  cause  great  surface  losses. 

Summing  up,  it  may  be  stated  that  the  application  of  the  different  means 
for  reducing  initial  condensation  resulted  in  the  common  use  of  the  two-cylinder 
compound  engine  for  condensing  service,  for  the  reason  that  the  low  pressure  part 
of  the  cycle  takes  place  in  the  saturated  region.  The  una-flow  principle,  however, 
permits  the  use  of  the  single  cylinder  single-stage  expansion  engine  for  this  ser- 
vice, since  the  uni-directional  flow  of  steam  eliminates  initial  condensation  despite 
the  fact  that  a  part  of  the  cycle  takes  place  in  the  saturated  region.  The  una- 
flow  principle  is  also  of  great  advantage  for  non-condensing  and  multi-stage 
engines. 


1* 


Ib.  The  Una-Flow  Arrangement  as  a  Means  for  Reducing 

Surface  Losses. 

The  una-flow  engine,  as  its  name  indicates,  utilizes  the  steam  energy  by  a 
um-directional  flow,  i.  e.  the  steam  passes  through  the  cylinder  always  in  the  same 
direction.  As  shown  in  Fig.  1,  the  steam  enters  the  cylinder  head  from  below,; 
heats  the  surface  of  the  latter,  and  then  enters  the  cylinder  through  the  inlet  valves 
located  in  the  top  portion  of  the  head.  Doing  useful  work,  the  steam  follows  the 


piston  and  after  having  expanded,  leaves  through  ports  at  the  opposite  end  of 
the  stroke,  i.  e.  in  the  middle  of  the  cylinder;  the  opening  and  closing  of  these 
ports  being  accomplished  by  the  piston  during  its  motion.  This  is  in  marked 
contrast  to  the  ordinary  or  counterflow  engine,  where  the  steam  enters  at  the 
end  of  the  cylinder,  follows  the  piston  during  the  working  stroke,  and,  returning 
with  the  piston,  leaves  at  the  cylinder  end.  The  result  of  this  kind  of  flow  is  an 
intensive  cooling  action  upon  the  clearance  surface,  the  exhaust  steam  being 


usually  wet  and  thus  an  excellent  heat  conductor.  The  consequence  is  increased 
initial  or  cylinder  condensation  during  the  following  admission.  The  una-flow 
principle  avoids  the  cooling  of  the  clearance  surfaces,  thereby  eliminating  initial 
condensation  to  such  an  extent  that  compounding  becomes  superfluous.  Una- 
flow  engines  may,  therefore,  be  built  with  a  single  cylinder  and  single-stage  expan- 
sion, and  yet  show  the  economy  of  compound  or  triple-expansion  engines. 

The  exhaust  ports  of  the  una-flow  cylinder  have  an  area  about  three  times 
as  large  as  can  be  realized  with  slide  or  poppet  valves,  with  a  consequent  complete 
pressure  equalization  between  cylinder  and  condenser  if  long  and  restricted  pas- 
sages between  them  are  avoided.  In  other  words,  if  the  condenser  is  placed  close 
to  the  cylinder  and  the  connection  is  of  ample  area,  then  complete  equalization  of 
pressures  is  assured.  In  order  to  get  a  clear  conception  of  the  magnitude  of  this 
port  area  one  has  to  consider  that  the  engine  piston  acts  as  a  piston  valve  and 
the  crank  as  eccentric,  while  the  cylinder  constitutes  the  valve  bushing.  The 
exhaust  lead  is  usually  taken  at  10%,  which  fixes  the  compression  at  90%. 

The  una-flow  exhaust  ports  do  away  with  separate  exhaust  valves  and  their 
leakage  loss,  their  additional  clearance  volume  and  surface,  as  well  as  the  neces- 
sary valve  gear.  The  elimination  of  exhaust  valves  is  therefore  an  additional 
advantage  of  the  una-flow  construction. 

Indicator  cards  of  una-flow  engines  show  adiabatic  lines  for  expansion  and 
compression. 

Adiabatic  expansion  results  in  considerable  moisture  even  with  highly  super- 
heated steam.  The  entropy  chart  shows  at  a  glance  that  with  an  initial  pressure 
of  12  at.  and  a  temperature  of  300°  C  an  expansion  to  0.8  at.  abs.  produces  7% 
moisture.  During  the  exhaust  period  this  expansion  continues  until  at  a  terminal 
'pressure  of  0.1  at.  abs.  the  moisture  amounts  to  17%.  On  account  of  the  un- 
dvoidable  heat  losses,  the  temperature  at  the  end  of  admission  will  be  somewhat 
less  than  the  above,  with  a  consequent  increase  in  the  final  moisture. 

The  extension  of  the  working  cycle  into  the  wet  region  rendered  compound 
engines  necessary,  the  high  pressure  cylinder  working  with  superheated,  and  the 
low  pressure  cylinder  with  saturated  steam.  The  una-flow  system  made  a  return 
to  the  single  stage  engine  possible  for  both  superheated  and  saturated  steam. 

Superheated  steam  is  a  very  effective  means  of  combating  initial  condensa- 
tion. The  combined  use  of  superheat  and  the  una-flow  construction  will  still  better 
conserve  the  heat  during  admission.  Expansion  will  therefore  start  at  a  higher 
temperature  and  terminate  with  less  moisture;  or  in  other  words,  better  economy 
will  result. 

The  head  jackets  have  the  effect  of  a  partial  regeneration  of  the  steam  during 
expansion  and  exhaust.  On  account  of  the  large  difference  of  temperature,  the 
large  heating  surface,  and  the  great  difference  in  the  specific  weights  of  the  live 
steam  and  the  exhaust  or  compression  steam,  an  intensive  heating  action  takes 
place  during  expansion,  exhaust,  and  compression.  This  affects  particularly  that 
part  of  the  steam  located  close  to  the  cylinder  head,  while  in  consequence  of  the 
adiabatic  expansion  that  part  following  the  piston  will  sustain  both  a  drop  in  tem- 
perature and  an  increase  in  moisture.  The  greater  part  of  the  moisture  produced 
will  accordingly  be  found  close  to  the  piston  head  with  a  progressive  dryness  and 


6 

an  increase  in  temperature  towards  the  cylinder  head.  This  moist  steam,  being 
close  to  the  exhaust  ports,  will  escape  first  when  they  are  uncovered,  while  that 
part  which  received  heat  from  the  head  jacket  is  trapped  by  the  returning  piston 
and  compressed  along  adiabatic  lines,  partly  as  saturated,  and  partly  as  super- 
heated steam,  but  mostly  the  latter,  because  during  the  early  part  of  compression 
heat  is  still  being  transmitted  to  it  from  the  head  jacket.  This  elimination  of 
liquid  condensate  avoids  its  deleterious  effect  of  heat  exchange  as  well  as  damage 
due  to  water  hammer. 

Tests  conducted  on  triple-expansion  engines  in  regard  to  the  action  of  steam 
jackets  have  shown  that  there  is  no  gain  in  high  pressure  cylinders,  very  little 
in  intermediate  cylinders,  but  a  large  gain  in  low  pressure  cylinders,  despite  the 
large  heat  losses  prevailing  in  cylinders  of  the  usual  type.  Owing  to  the  counter- 
flow  action,  a  great  part  of  the  jacket  heat  is  necessarily  carried  by  the  exhaust 
steam  into  the  condenser.  One  has  only  to  consider  that  at  the  time  of  exhaust 
valve  opening  a  considerable  amount  of  pressure  energy  is  available  to  exhaust 
the  steam  with  velocities  as  high  as  350  to  400  m.  per  sec.,  that  the  steam  with 
this  velocity  impinges  on  the  clearance  surfaces,  depositing  water,  and  that  on 
account  of  the  sudden  drop  in  pressure  the  heat  absorbed  by  them  during  admis- 
sion starts  an  intensive  re-evaporation  which  extracts  considerable  quantities  of 
heat  from  these  surfaces.  The  fact  that  the  latter  are  in  many  cases  jacketed 
presents  a  most  unfavorable  picture  of  the  very  inefficient  way  in  which  admission 
and  jacket  heat  are  utilized.  From  the  point  of  exhaust  valve  opening  up  to  the 
beginning  of  compression  the  jacket  heat  is  carried  into  the  condenser  in  the  most 
wasteful  fashion.  During  the  remaining  part  of  the  cycle,  heat  exchange  occurs 
under  more  unfavorable  conditions  and  at  lower  velocities,  but  notwithstanding 
the  immense  losses  of  jacket  heat  the  low  pressure  cylinder  derives  the  greatest 
benefit  from  steam  jackets.  This  may  be  explained  by  the  fact  that  the  low  pres- 
sure cylinder  has  the  greatest  temperature  differences,  the  greatest  heating  sur- 
faces, the  greatest  surface  losses,  and  a  very  favorable  density  ratio  of  jacket  to 
working  steam.  It  follows  that  una-flow  cylinders  necessarily  have  a  particularly 
energetic  heating  action,  since,  as  in  the  case  of  low  pressure  counterflow  cylinders, 
heating  takes  place  under  the  influence  of  the  full  temperature  difference,  the 
large  surfaces,  and  the  great  difference  in  density  of  jacket  and  working  steam. 
The  counterflow  of  steam,  with  its  great  losses,  is  replaced  by  the  uni-directional 
flow  where  no  jacket  heat  is  lost  to  the  exhaust.  As  shown  in  Fig.  2,  exhaust 
steam  in  a  una-flow  cylinder  never  passes  heated  surfaces.  The  layer  next  to  the 
cylinder  head,  at  the  very  worst,  approaches  the  exhaust  ports  without  leavipg 
the  cylinder,  and  therefore  hardly  any  jacket  heat  can  be  lost.  The  beneficial 
effect  of  steam  jackets  proved  to  e-xist  in  low  pressure  counterflow  cylinders  must 
therefore  be  in  evidence  in  a  high  degree  in  una-flow  cylinders,  since  a  loss  of 
jacket  heat  is  avoided. 

It  is  assumed,  of  course,  that  jacketing  is  limited  to  the  cylinder  head  and 
that  the  cylinder  is  unjacketed  (Fig.  1  and  2).  The  head  jacket  extends  to  the 
point  where  cut-off  normally  occurs  so  that  the  clearance  surfaces  of  the  first 
order  are  effectively  heated  from  the  outside,  while  the  highly  superheated  com- 
pression steam  prevents  cooling  from  the  inside. 


A  further  reduction  of  surface  losses  can  be  accomplished  by  enlarging  the 
jacket  surfaces  and  by  the  utmost  reduction  and  machining  of  clearance  surfaces. 
Even  though  the  clearance  surfaces  in  a  una-flow  engine  are  exposed  during  the 
whole  cycle,  they  are  not  subject  to  the  rush  of  exhaust  steam  passing  them,  nor 
to  re-evaporation;  and  they  therefore  suffer  little  cooling  on  account  of  the  com- 
parative tranquility  of  the  steam  molecules  adjacent  thereto,  and  furthermore  have 
the  benefit  of  both  the  jacket  and  compression  heat.  All  these  causes  combined 
with  the  una-flow  principle  and  una-flow  construction  produce  an  almost  adia- 
batic  action  of  the  clearance  surfaces. 


Fig.  2. 

The  una-flow  engine  fundamentally  avoids  the  thermal  mixup  characterizing 
the  counterflow  engine.  The  cylinder  is  composed  of  two  single  acting  cylin- 
ders with  the  exhaust  ends  in  common.  On  account  of  the  long  piston  the  stroke 
volumes  of  both  cylinder  ends  are  relatively  displaced  for  a  distance  equal  to  the 
length  of  the  piston.  The  two  inlet  ends  are  hot  and  remain  hot,  while  the  common 
exhaust  is  cold  and  remains  cold,  and  the  temperature  changes  gradually  from 
the  hot  inlet  to  the  cold  exhaust.  The  jacketing,  comprising  the  heating  cham- 
bers at  the  inlet  ends  and  the  cold  exhaust  belt  around  the  common  exhaust,  is 
in  perfect  accord  with  this. 

In  a  counterflow  cylinder  the  two  stroke  volumes  overlap.  The  exhaust  end 
of  one  side  more  or  less  reaches  to  the  admission  end  of  the  other,  according  to  the 
length  of  the  piston.  From  a  thermal  point  of  view  the  arrangement  is  obscure. 

The  central  exhaust  port  and  belt  in  the  middle  of  the  cylinder  effectively  cool 
that  part  of  it  at  which  the  piston  attains  its  highest  velocity.  This  favorable 


8 

action  is  further  augmented  by  the  omission  of  jackets  on  the  adjacent  part  of 
the  cylinder.  Furthermore,  the  piston  on  account  of  its  large  area  exerts  a  very 
low  unit  pressure.  The  cylinder  is  of  very  simple  form  and  can  be  kept  free  from 
badly  distributed  material,  thus  avoiding  local  heating  and  warping.  The  large 
bearing  surface  of  the  piston,  the  cooling  action  of  the  exhaust  belt,  as  well  as 
the  simplicity  of  the  cylinder,  which  precludes  the  possibility  of  warping  even 
at  the  highest  steam  temperatures,  render  the  use  of  a  tail  rod  unnecessary,  provided 
the  material  used  is  suitable,  the  design  correct,  and  proper  lubrication  supplied. 
(See  una-flow  locomotives  and  engines  built  by  Sulzer  Bros.)  The  piston  has  two 
sets  of  rings,  comprising  four  or  six  altogether.  During  the  period  of  high  pressures 
both  sets  are  in  action,  and  the  pressure  has  dropped  to  about  3  at.  when  one  set 
has  overrun  the  exhaust  ports.  The  many  una-flow  engines  in  operation  are  proof 
that  the  piston  is  not  the  cause  of  any  difficulties  even  with  the  highest  degrees 
of  superheat,  and  that  with  good  workmanship  a  piston  can  be  made  perfectly 
tight.  If,  however,  a  cylinder  should  become  scored,  its  simplicity  allows  it  to  be 
easily  replaced  without  great  expense. 

There  can  be  no  doubt  about  the  possibility  of  employing,  in  una-flow  engines 
of  the  kind  described,  steam  temperatures  far  in  excess  of  anything  used  at  present. 
Even  with  the  highest  initial  temperatures  the  cycle  extends  far  into  the  moist 
region,  thus  insuring  moderate  working  temperatures  for  cylinder  and  piston,  the 
highly  superheated  steam  being  limited  to  the  cylinder  head.  The  una-flow  engine, 
therefore,  opens  up  further  possibilities  of  development  in  the  utilization  of  higher' 
degrees  of  superheat.  It  is  all  the  more  suitable  for  this  because  the  superheat 
benefits  the  whole  cycle,  while  in  counterflow  compound  engines  the  high  pressure 
cylinder  receives  too  much  and  the  low  pressure  cylinder  too  little  superheat. 
This  feature  of  the  una-flow  engine  does  not,  however,  contradict  the  fact  that  it 
is  also  suitable  for  saturated  steam,  excellent  results  being  actually  obtained  with 
both  kinds.  The  una-flow  engine  has  the  uni-directional  flow,  the  hot  inlet  and 
the  cold  exhaust  in  common  with  the  steam  turbine.  From  a  thermal  point  of 
view  it  forms  the  missing  link  between  the  reciprocating  steam  engine  and  the 
steam  turbine. 

The  opinion  is  frequently  advanced  that,  in  regard  to  their  thermal  action, 
the  cylinder  head  and  piston  surfaces  of  a  una-flow  engine  are  merely  interchanged. 
This  leaves  out  of  consideration  the  fact  that  the  piston  surface  is  protected  against 
the  action  of  the  exhaust  by  a  cone  of  stagnating  steam.  The  existence  of  this 
phenomenon  has  been  frequently  proved  beyond  dispute  in  the  case  of  air  and 
water.  The  surface  corresponding  to  the  face  of  the  slide  valve  of  a  counterflow 
engine  is  located  in  the  una-flow  engine  at  the  circumference  of  the  cylinder.  The 
port  necessary  with  a  slide  valve,  together  with  the  clearance  and  clearance  sur- 
faces involved,  are  completely  avoided  in  the  una-flow  engine.  The  surfaces  of 
the  una-flow  exhaust  ports  are  outside  of  the  cylinder  and  therefore  have  no 
bearing  upon  the  thermal  conditions.  Piston  and  steam  have  a  different  velocity 
only  after  the  former  has  opened  the  exhaust  ports,  when  the  available  pressure 
energy  is  transformed  into  kinetic  energy.  The  steam,  however,  attains  its  full 
velocity  only  after  it  has  passed  the  edge  of  the  piston  and  therefore  the  cooling 
action  of  this  steam  upon  the  piston  can  only  be  small.  The  cooling  action  caused 


9 

by  the  low  velocity  of  approach  in  the  cylinder  now  remains  to  be  examined. 
Considering  the  exhaust  ports  as  nozzles  bounded  on  one  side  by  the  piston  edge, 
the  steam  at  this  narrowest  point  attains  a  mean  critical  velocity  of  about  410  m/sec. 
On  examining  cross-sections  ahead  of  this  smallest  section  or  throat,  very  moderate 
velocities  are  found.  Fig.  3  makes  these  conditions  clear  for  a  cylinder  of  600  mm 
bore  and  800  mm  stroke,  the  piston  being  assumed  to  have  overrun  the  ports  for 
a  distance  of  20  mm.  Calculation  of  the  velocity  at  this  narrowest  point  shows 
it  to  be  410  m/sec.,  while  at  a  distance  of  15  mm  ahead  of  this  point  it  is  only 
71  m/sec.;  at  40  m  it  is  36  m/sec.;  at  85  m  it  is  20  m/sec.,  and  at  130  m  it 
is  15  m/sec.  It  should  be  noted  that  in  considering  the  various  cross-sections 


a  reduction  of  area  due  to  the  bridges  has  been  assumed  at  the  narrowest  section 
(throat),  and  this  does  not  apply  to  the  cross-sections  of  approach.  As  the  piston 
progressively  uncovers  the  exhaust  ports  the  proportions  of  the  cross-sections  are 
changed  in  such  a  way  that  the  velocity  of  approach  is  increased.  At  the  same 
time  a  reduction  of  pressure  occurs,  the  effect  of  which  is  to  reduce  the  velocity 
of  approach,  beginning  at  the  point  at  which  the  critical  pressure  is  reached. 
Most  indicator  cards  of  una-flow  engines  show  clearly  that  when  the  piston  reaches 
the  dead  center  the  pressure  inside  the  cylinder  has  dropped  to  the  back  pressure; 
or  in  other  words,  the  greater  part  of  the  steam  has  in  this  position  been  exhausted. 
Therefore,  owing  to  the  short  duration  and  low  intensity  of  the  flow  of  exhaust 
steam  along  the  piston  surface  and  the  small  harmful  exhaust  surfaces,  the  resulting 
cooling  action  is  small.  The  whole  cylinder  section  acts  as  an  approach  to  the  exhaust 
nozzles.  A  further  protection  is  given  by  the  layer  of  stagnating  steam,  and  there 
is  finally  a  very  intense  heating  action  during  compression  and  admission.  The 


10 

heating  of  the  piston  surface  by  the  hot  live  steam  is  so  effective  that  this  surface 
acts  almost  adiabatically  during  the  following  exhaust  period.  The  very  favorable 
steam  consumption  figures  obtained  with  this  type  of  engine  are  further  proof 
that  a  simple  interchange  of  the  cylinder  head  and  piston  surfaces,  in  regard  to 
their  thermal  behaviour,  is  out  of  the  question.  Such  favorable  economy  can  only 
result  if  the  piston  and  its  cooling  action  is  negligible.  This  is  further  confirmed 
by  tests  of  Prof.  Nagel.  A  test  with  saturated  steam  of  184°  C,  and  a  cut-off 
at  12%,  showed  that  the  temperature  of  the  piston  surface  at  a  point  near  its 
circumference  was  about  164.5°  G.  The  total  fluctuation  at  this  point  was  only 
1.3°  G.  This  surprisingly  high  temperature,  and  moreover  the  small  fluctuation, 
justifies  a  very  favorable  conclusion.  These  figures  should  be  still  better  towards 
the  center  of  the  piston  surface. 

The  thermal,  constructional,  and  operative  advantages  of  this  type  of  prime 
mover  are  such  that  in  continuous  operation  the  economies  of  compound  and 
triple-expansion  engines  can  be  obtained  with  both  saturated  and  superheated 
steam. 

Jacketing  of  the  Cylinder. 

The  firm  of  Sulzer  Brothers,  of  Winterthur,  Switzerland,  constructed  an  ex- 
perimental engine  of  the  una-flow  type  after  such  engines  had  been  put  on  the 
market  by  the  Erste  B runner  Maschinenfabrik  Gesellschaft,  who  were  the  first 
to  take  up  the  manufacture  of  una-flow  engines  on  the  author's  recommendation. 


Fig.  5. 

Sulzer  Bros,  entrusted  the  author  with  the  design  of  the  first  una-flow  cylinder 
of  600  mm  bore,  800  mm  stroke,  and  155  r.  p.  m.  All  later  Sulzer  engines  are 
built  with  only  slight  changes  from  this  design,  which  is  shown  in  Figs.  4  and  5. 
According  to  Sulzer  Brothers'  usual  practice  the  engine  was  put  on  the  testing 
floor  and  a  great  number  of  tests  were  carried  out  with  the  object  of  observing 


11 


its  performance  and  determining  its  economy  under  the  most  varying  conditions. 
An  important  part  in  the  program  was  the  study  of  the  effect  of  the  jackets.  For 
this  reason  the  author  incorporated  in  this  design  not  only  head  jackets,  through 
which  the  live  steam  had  to  pass  before  entering  the  cylinder,  but  also  jackets 
at  the  ends  of  the  cylinder  barrel,  which  were  separated  by  a  neutral  zone  from 
the  exhaust  belt.  These  cylinder  jackets  could  be  shut  off  separately.  During 
the  tests  the  head  jackets  were  necessarily  always  in  operation,  but  the  cylinder 
jackets  were  either  in  service  or  shut  off,  as  indicated  in  Fig.  6  by  the  words 
,,with  jacket"  and  ,, without  jacket".  The  tests  proved  that  the  effect  of  the 
cylinder  jacket  decreases  with  increasing  steam  temperature.  With  saturated 
steam  the  difference  was  almost 
1  kg/I  HP-hour  in  favor  of  cylinder 
jackets,  while  it  was  barely  0.5  with 
steam  of  265°  C  and  only  0.2  kg  with 
steam  of  325°  C.  All  figures  refer- 
red to  the  most  economical  M.  E.  P. 
The  steam  pressure  was  9.2  at. 
gage  and  the  vacuum  66  cm. 

For  the  point  of  best  operating 
economy,  i.  e.,  an  M.E.P.  of  about 
3  kg/sqcm,  these  differences  change 
in  such  a  way  that  a  small  increase 
results  when  running  with  cylinder 
jackets  and  with  a  steam  tempe- 
rature of  325°  G,  while  steam  of 
265°  G  shows  a  small  decrease  in 
steam  consumption. 

For  a  steam  temperature   of 


kg/cm* 


325°  C  the  point  of  equality  of 
steam  consumption,  when  operating  both  with  and  without  cylinder  jackets,  is 
found  to  correspond  to  an  M.E.P.  of  2.5  kg/sqcm.  The  corresponding  point  for 
steam  of  265°  G  occurs  at  an  M.E.P.  of  3.4  kg/sqcm.  For  saturated  steam  this 
point  moves  towards  a  still  higher  M.E.P. 

This  explains  the  customary  omission  of  cylinder  jackets  for  superheated  steam. 

It  is  also  noteworthy  that  the  best  economy  with  steam  of  325°  G  very  clo- 
sely approaches  the  value  of  4  kg/I  HP-hour.  The  results  for  saturated  steam, 
especially  with  cylinder  jackets,  are  extremely  favorable.  It  must  be  considered, 
however,  that  the  saturated  steam  had  a  very  slight  degree  of  superheat  in  order 
to  make  sure  that  it  was  actually  dry.  It  should  further  be  noted  that  this  engine 
was  well  designed  and  well  built.  The  clearance  volume  and  clearance  surface 
(the  latter  being  machined)  were  moderate,  and  the  whole  engine  was  built  with 
the  high  precision  usual  to  Sulzer  Brothers'  shop  practice.  The  measurements 
were  made  by  means  of  a  surface*  condenser,  which,  as  is  well  known,  gives  slightly 
lower  but  more  accurate  results  than  boiler  feed  water  measurements. 

Comparing  these  curves  with  those  of  compound  or  triple-expansion  engines, 
it  will  be  found  that  the  steam  consumption  of  the  una-flow  engine  is  influenced 


12 


Fig.  7. 


by  the  load  to  a  much  smaller  degree.  This  is  shown  particularly  by  the  curves 
marked  ,, without  jackets'"',  where  there  is  hardly  any  change  in  the  steam  con- 
sumption between  mean  effective  pressures  of  1  and  3  kg/sqcm,  especially  with 
high  superheat.  Even  with  saturated  steam  little  change  is  noticeable  between 
mean  effective  pressures  of  1  and  2,4  kg/sqcm. 

The  curves  of  Fig.  6  justify  the  following  conclusions  in  regard  to  cylinder 
jacketing: 

For  highly  superheated  steam  (300°  C  and  more)    and  high  mean  effective 
pressures  cylinder  jacketing  is  useless  (Fig.  7),  and  has  a  deleterious  effect  even 

at  as  low  an  M.E.P.  as  3  kg/sqcm.  For 
low  mean  effective  pressures  cylinder 
jackets  may  yet  be  expected  to  yield  a  small 
gain.  For  moderate  steam  temperatures  of 
about  250° G  a  short  cylinder  jacket  as  shown 
in  Fig.  8  is  advisable.  It  will  slightly  im- 
prove the  economy  even  for  a  high  M.E.P. 
and  produce  considerable  gain  for  a  low 
M.E.P.  With  saturated  steam  or  low  de- 
grees of  superheat  a  cylinder  jacket  separa- 
ted only  by  a  narrow  zone  from  the  exhaust 
belt  (Fig.  9)  should  be  used  under  all  circum- 
stances. Head  jackets  are  essential  in  all  cases. 
The  separation  of  cylinder  jacket  and 
exhaust  belt  is  advisable  in  order  to  avoid 
unnecessary  loss  of  jacket  heat  and  to  pro- 
vide more  favorable  operating  conditions 
for  the  piston,  especially  when  using  super- 
heated steam.  The  center  part  of  the  cy- 
linder, where  the  piston  attains  its  highest 
speed,  has  the  lowest  temperature.  Ex- 
cellent results  can  be  secured  with  self- 
supporting  pistons,  even  with  very  high 
temperatures  of  superheat,  if  the  designer 
pays  attention  to  these  thermal  conditions 
by  providing  the  piston  with  bearing  sur- 
faces at  its  center  only,  leaving  its  extremities  to  project  in  the  manner  of  a 
plunger  towards  both  ends  of  the  cylinder  without  actually  touching  the  walls. 
The  head  jackets  do  not  impair  the  operation  of  the  piston,  since  no  rubbing 
surfaces  are  in  contact  with  the  former.  Their  heating  action  is  very  effective 
because  they  are  continuously  in  contact  with  the  working  steam ;  they  heat  harmful 
surfaces  of  the  first  order,  and  with  proper  lubrication  the  transmission  of  heat 
from  them  is  not  impeded  by  an  oil  film  (Fig.  7).  There  is  practically  no  loss  of 
jacket  heat  to  the  exhaust.  Conditions  are  not  so  good  in  Fig.  8  and  still  worse 
in  Fig.  9.  In  these  constructions  the  amount  of  jacket  heat  lost  to  the  exhaust 
increases  more  and  more  because  the  exhaust  steam  to  a  certain  extent  flows  past 
heated  surfaces,  although  these  may  be  partly  protected  by  an  oil  film. 


Fig.  8. 


13 


2  a.  Influence  of  the  Clearance  Volume  upon 
the  theoretical  Steam  Consumption  (Volume  Loss). 

(The  Una- Flow  System  as  a  Means  for  Reducing  the  Volume  Loss). 

A  certain  amount  of  steam  admitted  per  stroke  into  a  cylinder  with  clearance 
will  produce  an  indicator  card  of  less  area  than  in  an  ideal  cylinder  without  clea- 
rance. The  difference  in  area  may  be  termed  volume  loss.  This  volume  loss  is 
represented  in  Figs.  1  to  4  for  different  conditions.  The  diagrams  corresponding 
to  the  cylinder  with  clearance  are  drawn  in  heavy  lines,  while  those  for  the  ideal 
cylinder  are  shown  dashed. 

Volume  losses  can  be  expressed  absolutely  or  relatively.  In  most  cases  it  is 
convenient  to  figure  the  volume  loss  in  per  cent  of  the  engine  output. 


u-^— H 


Fig.   1. 


A  comparison  of  the  two  areas  AO PG  and  ABPG  in  Fig.  1  shows  area  BOP 
to  be  a  loss.   The  parts  of  the  diagrams  lying  below  the  line  G  P  show  a  loss  of  the 
area  GES  and  a  gain  of  area  PCV  =  GFQ  T.    By  subtracting  the  latter,  the 
resultant  loss  is  represented  by  the  area  TQFES. 

In  Fig.  2  the  admission  has  been  lengthened  until  the  points  F  and  .P  of 
Fig.  1  coincide  with  the  beginning  and  ending  of  the  diagram.  Consequently  the 
shaded  areas  BOP  and  GES  represent  the  loss  for  the  diagram  with  clearance. 

In  Fig.  3,  with  still  longer  admission,  a  comparison  of  the  diagrams  shows 
the  loss  to  be  equal  to  the  shaded  areas  BOVC  and  GHS. 


14 


The  reduction  in  diagram  area  produces  a  corresponding  increase  of  the  loss 

due  to  incomplete  expansion.    A  special  case  is  shown  in  Fig.  4,  which  represents 

•  a  diagram  without  volume  loss.    This  may  occur  for  instance  in  high  pressure 

cylinders  of  compound  engines  when  expansion  is  carried  to  the  back  pressure 

and   compression  to  the  initial   pressure.    Although  the  diagram  has  no   direct 

volume  loss,  the  stroke  volume  of  the  cy- 
linder with  clearance  must  be  increased 
from  F!  to  F2,  with  a  consequent  increase 
in  the  surface  loss. 


\d     O 


\r 


Fig.  3. 


Fig.   4. 


The  diagrams  clearly  disclose  the  fact  that  the  compression  is  a  means  of 
reducing  the  volume  losses,  since  they  would  be  essentially  larger  without  com- 
pression. 

It  is  also  clear  that  for  a  constant  admitted  steam  volume  cp  the  point  of 
cut-off  will  vary  with  different  lengths  of  compressions.  There  will  therefore  be 

a  certain  position  of  9?  for  which  the  diagram 
area  produced  will  be  a  maximum  and  the 
corresponding  volume  loss  a  minimum. 

\    X^  Fig.  5.  a)  Determination   of  the   best  position 

\        \  of  9?,  if  pv  =  constant,   and  /?1?  /?2,  cp  and 

^  SQ  are  known  (see  Fig.  5). 


Diagram  areae  F  =  ABEDRF  = 
=  h  +  h-h-h=ABHG  + 
BEKH  —  FRJG  —  RDKJ. 


\\ 


15 


The  admitted  steam  volume  <p  ---  VE--  Vc   and  Vc  =  —  •  VK  and  VK  - 


VH  y 

Therefore:    /1==  p1  (VE  —  SQ);     f2  =  §pdv  =  VEpL  loge  —  ^-; 


I/ 

loge      ~^-    =  (VE—(f>)piloge 

=  Pi(VE  —  S0)  +  VE /?!  loge  -jr~—  (VE  —  <P)  Pi  loge 


-VH-PZ  +  (VE  —  <p)  •  Pi- 

Required  is  the  maximum  diagram  area  for  <p  =  constant.    The  independent 
variable  is   VE-     Therefore: 

VH 

dF  VH  VE*  VE  —  <p     Pl 

-jy-  ->i  +  Pi  loge  ~y  --  VE-p1-v-    -  pl  loge  --  *- 

«  y  E  V  E  V  H.  *->0  Pz 


B  —  99      Dj         A  H  £  —  <p 

—  Pl   log*    --  C-^'  =0»    OF  "  ^c—  ~ 


or  VE=  —  ±.\-L-  -f  VH-SQ-      ......         (I). 

Since  (VE  —  <p)—     =  V K,     therefore  jr^~-/-    or  ±±  =  ±±     .     .       (II). 
Pz  VE         <J0  pe        p2 

This  result  is  graphically  represented  in  Fig.  5.  A  straight  line  is  drawn 
through  D  and  B,  and  intersects  the  vertical  axis  at  L.  Another  line  is  drawn 
through  L  and  A,  intersecting  the  back  pressure  line  at  /?,  and  this  point  deter- 
mines the  best  compression  for  the  assumed  values  of  VE  and  S0.  Another  method 
consists  in  finding  the  intersection  with  the  horizontal  axis  of  a  line  through  A 
and  E,  and  drawing  a  line  from  this  intersection  through  the  point  D  to  cut  the 
vertical  line  through  A  at  the  point  F,  which  latter  thus  determines  the  best  ter- 
minal compression  pressure. 

The  method  of  Fig.  5  cannot,  however,  be  used  to  find  Vk  if  <p  is  given. 
In  this  case  equation  I  may  be  used  to  find  VE,  and  when  this  quantity  is 
known  Vk  can  be  found  either  by  equation  II  or  the  graphic  method  shown 
in  Fig.  5. 

From  Fig.  5  it  will  be  seen  that  early  cut-offs  are  associated  with  long  com- 
pressions and  vice  versa.  In  case  of  a  cut-off  at  100%  the  intersection  of  the  line 


16 


DB  with  the  vertical  axis  moves  to  infinity.  Consequently  the  intersection  of 
the  line  A  R  with  this  axis  must  be  at  infinity ;  in  other  words  the  line  is  parallel 
to  the  axis  and  thus  gives  a  length  of  compression  equal  to  zero.  For  a  cut-off 
equal  to  0%,  the  two  lines  AR  and  BD  coincide  and  the  compression  therefore 
is  100%. 

Fig.  6  represents  the  case  in  which  the  terminal  expansion  pressure  equals 
the  back  pressure.  The  corresponding  compression  reaches  the  initial  pressure, 
this  being  a  condition  which  remains  true  also  for  polytropic  lines. 

At  a  smaller  cut-off  than  a  in  Fig.  6  the  expansion  and  compression  lines 
both  form  loops,  since  the  terminal  expansion  pressure  falls  below  the  back 
pressure  and  the  compression  exceeds  the  initial  pressure.  A  loop  in  the  ex- 
pansion line,  therefore,  corresponds  to  a  loop  in  the  compression  line.  Such  a 
diagram  laid  out  according  to  Fig.  5  would  be  correct  from  the  point  of  view 
of  volume  loss. 

Large  clearance  volumes  are  accompanied  by  large  volume  losses,  and  on 
the  basis  of  these  great  losses  a  change  in  the  length  of  compression  may  produce 

a  considerable  gain.  Small  clearance  volumes 
permit  the  use  of  constant  compression, 
since  a  change  in  the  length  of  compression 
could  result  only  in  a  small  gain  in  view  of 
the  slight  volume  loss,  and  with  many  types 
of  valve  gear  such  a  change  would  give 
inadmissibly  high  compression  pressures.  Non- 
condensing  engines,  such  as  locomotives  for 
instance,  which  have  about  12%  clearance, 
are  usually  fitted  with  gears  which  vary 
the  length  of  compression  inversely  with  the 
cut-off.  A  link  valve  gear,  aside  from  the 
large  clearance  volume  caused  by  its  use, 
gives  a  qualitatively  correct  change  in  the 
length  of  compression  and  also,  as  will  be  proved  later,  a  change  which  is 
also  quantitatively  correct.  Similar  conditions  prevail  in  a  single  valve  non- 
condensing  engine  having  about  12%  clearance  in  which  the  shaft  governor 
alters  the  throw  and  angle  of  advance  of  the  driving  eccentric. 

The  large  clearance  necessitated  by  both  these  types  of  valve  gear  is  a  great 
disadvantage,  which  is  only  partly  remedied  by  a  correct  variation  in  the  length 
of  compression. 

In  the  case  of  condensing  engines  a  variation  of  the  length  of  compression 
gives  very  little  improvement,  as  will  be  shown  later.  Constant  compression  is 
allowable  and  long  compression  desirable. 

The  use  of  constant  compression  for  small  clearances  is  in  accord  with  the  case 
of  zero  clearance  volume,  since,  in  accordance  with  Fig.  6,  it  requires  a  constant 
length  of  compression  equal  to  zero  for  a  cut-off  at  any  point  of  the  stroke. 

b)  Determination  of  the  best  position  of  <p  when  pvn  =  constant  and  pA,  p2, 
99  and  S0  are  known  (see  Fig.  5). 


Fig.  6. 


17 


i—n 


Pi 


—  np2 


i  — n 

~n  —  So1"")   or  with    VK  - 
j_ 

/T7  ~\  Pi          C    l-n  /I 


n 

(VE-<p)  £- 

\Pzl 


i—n 


Pz 


(VE-V). 


These  areas  as  previously  combined  give: 


i—n 


i—n 


i—n  \p2 


dF 
For  a  minimum  value  of  jP,  =  0. 


dF 
dVE 


P2 


i—n  ' 


i  —  n        i  —  n   \pz 


i\ 

TT~        ~T 


- 


=  0 


4. 


__  .p 
i-n    Pz\p 


or: 


H 


n-l 


Fig.  7. 


The  una-flow  steam  engine. 


The  quantities  F£,  F^  and  e  giving  the 
maximum  diagram  area  for  constant  value  of  <p 
are  obtained  from  equations  I,  II  and  III. 

Plotting  the  calculated  values  of  e,  VK  and 
the  corresponding  mean  effective  pressures  pt-  in 
a  diagram  (Fig.  7)  with  99  as  abscissae  and  VE,, 
VK   and  pi  as    ordinates,   the   best 
combination  of  9?,  VE  and  V  K  may 
be  read  off  for  a  given  value  of  p{. 
A  different  problem  is  presented 
by  the  determination  of  that  length 
of  compression  K  which  results   in 
the  lowest   steam  consumption   for 
a  given  constant  quantity  e  (Fig.  8). 


18 

a)  Assuming  pv  =  constant. 

Equations  for  /1?  /2,  /3,  /4  are  as  before.    The  smallest  steam  consumption 

C  '=  ^  is  to  be  found;   or  for  simplicity,   the  maximum  value   of  the   fraction 
F 


Hence  0= 


dO 
The  independent  variable  is   V  K,  therefore  -7^7-  =  0 


Pi, 

=0 


Pi. 

0+M^E-^o) 


Pi  Pi 

^VK\ogeVK  +  P^VKlogeS0- 
Pi  Pi  Pi 


or: 


.     (IV). 


b)  pyn  =  constant. 
As  previously, 


for  smallest  steam  consumption 

=  0,  therefore 


dV 


K 


^^il(W^^ 

XPi/    H \i  —  n  l     i  —  n ]  _    . 


19 


Again, 


v-«-F^-«)+^(^i-"-v-n)+p2(^-^)l|-Hn! 
1  At  j  i  \pi/  \ 


Pit    I 


_n 

n-lp 


n-l 


Equations  III,  IV  and  V  can  be  solved  only  by  trial  and  error.  Furthermore, 
the  expansion  and  compression  lines  especially  of  superheated  steam,  do  not  clo- 
sely follow  the  law  pvn  =  constant;  the  following  method  based  on  adiabatic 
expansion  and  compression,  and  on  the  use  of  the  Mollier  chart  is  therefore  pre- 
ferable because  it  permits  of  the  exact  determination  of  the  steam  consumption 
for  given  values  of  the  quantities  p^  pz,  S0  VE  and  VK. 

It  is  assumed  that  the  following  conditions  hold  true:  Adiabatic  expansion 
and  compression,  quality  of  steam  at  beginning  of  compression  equal  to  quality 
obtained  by  expanding  adiabatically  to  the  back  pressure,  cylinder  stroke  vo- 
lume =  1  sqm  X  1  m  =  1  cbm;  gain  of  heat  due  to  bringing  compression  steam 
up  to  initial  pressure  neglected. 

The  following  symbols  are  used: 

A  ie  Change  of  total  heat  during  expansion  in  cal/kg. 
Aie        ,,        „       „        ,,          „       compression  in  cal/kg. 


S0  Clearance  volume  in  ° 


J/o- 


P!,  V1  Initial  pressure  kg/sqcm  and  specific  volume  cbm/kg. 
pe,  Ve  Terminal  expansion  pressure  kg/sqcm  and  specific  volume  cbm/kg. 
/?2,  V2  Back  pressure  kg/sqcm  and  specific  volume  cbm/kg. 
pk  vk  Terminal  compression  pressure  kg/sqcm  and  specific  volume  cbm/kg. 
e  Length  of  admission  in  %. 
k  Length  of  compression  in  %. 

vc  Volume  of  compression  steam  reduced  to  initial  pressure  in  percent 
of  stroke. 

The  area  of  the  indicator  card  is  sub- 
divided into  the  following  four  parts  (Fig.  8) : 


/3  =  F'FED' ;        /4  =  A'AFF'. 

Determination  of  these  separate  areas: 

a)  Area  f±:  The  work  for  1  kg  of  steam 

expanding   adiabatically  from   p1   to   pe  is 

427  •  A  ie  mkg.    In  a  cylinder  of  1  cbm  stroke 

volume  the  weight  of  the  working  steam  is 


100 


kg, 


Fig.  8. 


20 


therefore  the  work  represented  by  the  area  fl  is  equal  to 


b)  Area  /2:  For  a  cylinder  having  1  cbm  stroke  volume  the  work  is  equal  to 

L2  =  10  000  (p.  -  p2)  1°°1+tSo  mkg. 

c)  Area  /3:  Again,  the  work  for  1  kg  steam  =  427  •  A  ic  mkg.    In  a  cylinder 
of  the  assumed  stroke  volume  the  weight  of  compression  steam  is 

~TOO~~'^kg' 
Therefore,  the  work  corresponding  to  area  /3  is 

•  —  mkg- 


d)  Area  /4  :  For  the  same  stroke  volume, 


e)  The  area  of  the  indicator  card  is:  F  =  fl  +  /2  —  /3  —  /4,  and  the  indi- 
cated work  is:  Lt  =  Lj  +  L.,  —  Ls  —  L4  mkg. 


f)  The  mean  effective  pressure:  pt  = 


kg/sqcm. 


g)  Steam  consumption:  The  admitted  weight  of  live  steam 

-i^r~'^kg 

The  corresponding  work  is  L,  mkg.    The  work  for  one    H.  P.   hour   being 
=  75  •  60  •  60  =  270000  mkg,  the  steam  consumption  is 

270000    e  +  S0-Vc    i  ^  UD  ^_  (yi> 


The  quantities  Fc,  pe,  Zl  ie,  ^  ic  and  pfc  can 
be  obtained  directly  from  the  Mollier  chart. 

The  values  plotted  in  the  following  diagrams 
were  obtained  by  this  method. 

In  Figs.  10,  11,  12,  13,  14  and  15  the  steam 
consumption  for  an  initial  pressure  of  14  at. 
abs.,  1  at.  abs.  back  pressure,  superheated  steam 
of  300°  G  and  clearance  volumes  of  5,  8  and 
11%,  has  been  plotted  against  the  length  of 
compression  A;,  for  various  values  of  constant 
values  e  and  constant  M.  E.P.  The  construction 
of  auxiliary  diagrams  (Fig.  9)  is  recommended 
for  determining  the  curves  of  M.  E.P.,  the 
abscissae  representing  M.E.P.,  and  the  ordi- 
nates  the  steam  consumption.  These  curves  are 

plotted  for  constant  compression,  and  a  vertical  line  intersecting  them  determines 
the  values  of  the  steam  consumption  for  a  given  constant  mean  effective  pressure. 


Meow  e£  pir 

Fig.   9. 


21 

The  great  influence  exerted  by  the  compression  upon  the  steam  consumption 
is  shown  in  Figs.  10,  11,  12,  13,  14  and  15.    As  indicated  by  previous  results,  the 


Fig.  10.     Saturated  steam  14  at.  abs. 
Noncondensing,   Clearance  space  5°/0. 


Fig.  11.     Saturated  steam  14  at  abs. 
Noncondensing,   Clearance  space  8°/0. 


best  economy  is  obtained  with  long  compressions  for  early  cut-offs  e  and  small 
mean  effective  pressures  pt  and  vice  versa.  It  will  also  be  noticed  that  link  valve 
gears  and  valve  gears  controlled  by  shaft  governors  acting  upon  both  inlet  and 


22 


exhaust  have   approximately  correct  variation  of  compression   (Fig.  14).    Valve 
gears  operating  with  fixed  compression  can  rightly  be  used  in  connection  with 

small  clearance  volumes.  Fig.  14  is 
especially  interesting,  since  it  also  con- 
tains the  curve  of  compression  obtained 
with  the  standard  Walschaert  gear  of 
the  German  State  Railways  for  an  ex- 
haust lap  of  —  3  %  mm.  It  is  surprising 
to.  find  that  this  curve  agrees  closely 
with  the  calculated  minimum  values 
of  the  compression  for  constant  M.  E.P. 
The  shortest  cut-off  of  the  link  valve 
gear  necessitates  a  large  clearance  vo- 
lume, the  bad  effect  of  which  is  only 
partly  neutralized  by  a  correct  variation 
of  compression.  It  would  be  more  im- 
portant, however,  considerably  to  reduce 
the  clearance  volume  required  by  this 
type  of  gear,  since  the  bad  effect  of 
large  clearance  volume  far  outweighs 
the  correcting  influence  of  the  com- 
pression. 

The  best  steam  consumption  for 
constant  mean  effective  pressure  pt  on 
the  one  hand,  and  constant  cut-off  e 
on  the  other,  occur  at  different  lengths 
of  compression.  If,  as  it  must  be,  pt  is 
considered  the  governing  variable,  then 
shorter  compressions  are  arrived  at.  The 
smaller  the  clearance  volume  is,  the 
shorter  will  be  the  compressions  at 
which  both  these  minima  occur. 

Figs.  16  and  17  explain  the  different 
sets  of  curves  more  clearly.     In  Fig.  17 
are  repeated  two  curves  from  Fig.  15, 
one  being  a  constant  M. E.P.  curve  for 
Pi  =  10  at.,   and  the  other  a  constant 
cut-off  curve  for  e  =  50%.  The  diagrams 
in  Fig.  16  marked  A  (9%  compression), 
and  B    (76%    compression),    have   the 
same  theoretical  steam  consumption  of 
8   kg/HP-hour,     the    cut-off    being    in 
both  cases  at  50%  of  the  stroke.    Dia- 
gram C,   also  with  a  cut-off  at  50%  but  with  47%  length  of  compression,   has 
a  steam  consumption  of  only  7.85  kg,   while  still  another  diagram   D,    having 
the  same  M. E.P.    (10  at.)   as  diagram  C,    but  with   a  cut-off    at  44%    and   a 


Fig.  12.     Saturated  steam  14  at.   abs. 
Noncondensing,   Clearance  space  11  °/0. 


23 

compression  of  19%,  only  has  a  steam  consumption  of  7.6  kg.  This  last 
diagram  is  represented  by  the  lowest  point  of  the  curve  for  constant  M.E.P. 
(Fig.  17).  At  this  point  D,  however,  the  constant  M.E.P.  curve  is  intersected 


Fig.  13.  Superheated  steam  300°  C,  14  at.  abs.       Fig.  14.   Superheated  steam  300°  C,  14  at.  abs. 
Noncondensing,   Clearance  space  5°/0-  Noncondensing,  Clearance  space  8°/0. 

by  a  curve  of  constant  cut-off  at  44%.  The  lowest  point  of  the  latter  in  turn  is 
intersected  by  another  M.E.P.  curve  which  also  has  a  minimum.  By  following 
from  minimum  to  minimum  along  the  M.E.P.  and  cut-off  curves  a  point  is 


24 


10 


finally  reached  at  which  the  .two  minima  coincide.  This  point  represents  a  dia- 
gram with  complete  expansion  and  with  compression  to  the  initial  pressure, a  dia- 
gram which,  as  previously  proved  (Fig.  4),  has  no  volume  loss  and  gives  the  lowest 
theoretically  possible  steam  consumption  for  the  assumed  range  of  pressures. 

Similar  curves  are  shown  in 
Fig.  18  for  condensing  operation. 
They  refer  to  superheated  steam  of 
13  at.  abs.,  a  steam  temperature  of 
300°  C,  a  back  pressure  of  0.08  at., 
and  a  clearance  volume  of  2%.  In 
this  diagram  also  the  M.E.P.  curves 
essentially  determine  the  best  com- 
pression. 

For  the  M.E.P.  of  2  to  3  at. 
ordinarily  used  it  is  evident  that  the 
best  compression  approximates  90% 
at  this  low  back  pressure,  but  even 
for  considerably  higher  mean  effec- 
tive pressures  the  difference  in  steam 
consumption  between  90%  and  the 
best  compression  is  negligible.  The, 
difference  gradually  disappears  as  the 
back  pressure  approaches  the  absolute 
vacuum,  since  in  this  case  com- 
pression naturally  would  have  no  in- 
fluence whatever  upon  the  steam 
consumption.  Nevertheless,  for  2% 
clearance,  superheated  steam  of  13  at. 
abs.  having  a  temperature  of  300° G, 
an  M.E.P.  of  2,8  at.,  and  a  back 
pressure  of  0.044  at.  abs.,  the  best 
compression  is  90%.  These  may  be 
considered  average  conditions  for  con- 
densing una-flow  engines. 

This  proves  conclusively  that  the 
long  compression  of  the  una-flow  en- 
gine is  in  no  way  a  necessary  evil 
accompanying  the  use  of  piston-con- 
trolled exhaust  ports. 

The  flatness  of  the  M.  E.  P.  curves 
also  indicates  that  it  is  permissible  to 
keep  the  compression  constant  in  the 


20          W         60        <30        100 
Compress/on  //?  %* 

Fig.  15.     Superheated  steam  300°  C,  14  at.   abs. 
Noncondensing,   Clearance  space  11°/0. 


above  case,  which  is  a  further  argument  for  the  correctness  of  the  una-flow  system. 
The,  long  and  constant  compression  of  90%  of  the  condensing  una-flow  engine 
is  therefore  correct  and  admissible. 


25 


Many  authors  discuss  the  ''high  compression"  of  the  una-flow  engine  in  the 
sense  of  its  being  unavoidable  or  undesirable.    "High  compression"  is  evidently 

confused  with  "long  compression".  A 
compression  line  may  be  long  with  low 
terminal  pressure,  or  short  with  high  ter- 
minal pressure.  Generally  speaking,  ter- 
minal compression  pressures  are  too  low 
in  the  majority  fof  condensing  una-flow 


Fig.   16. 


5,0 


X 

• 


I 


3,5 


«5 


'eat 


^» 


20 


ffff        00 


Fig.  17.   Superheated  steam  300°  C,  14  at. 
abs.  Noncondensing,  Clearance  space  11  °/0. 


Fig.  18.    Superheated  steam  300°  C,  14  at  abs. 
Condensing  (0.08  at.  abs.)   Clearance  space  2°/o- 


engines,  and  should  be  considerably  higher  according  to  paragraph  9  of  the  summary. 
Steam  consumption  and  compression  curves  Jfor  saturated  steam,  correspond- 
ing to  those  for  superheated  steam  shown  in  Fig.  18,  would  show  only  a  slight 
deviation  from  the  latter,  with  the  effect  that  the  best  compressions  are  slightly 


26 


shorter  throughout.    The  assumption  should,  however,  be  remembered  that  ex- 
pansion and  compression  are  adiabatic  and  that  the  compression  is  thought  to 


cf          18         Iff 
3Md//C/?er  flat/m  in 


Fig.  19.     Saturated  steam  14  at.  abs. 
Noncondensing. 


C/ect  ranee  s/icrce 


Fig.  20.   Superheated  steam  300°  C, 
14  at.  abs.   Noncondensing. 


begin  with  the  quality  of  »team  resulting  from  continued  adiabatic  expansion 
during  exhaust. 


Jacketing  of  the  cylinder,  especially  in  una-flow  engines,  requires  shorter  com- 
pressions, since  the  effect  of  the  jacket  is  to  increase  the  temperature  of  the  resi- 
dual steam,  thereby  superheating  it  at  an  earlier  stage  and  thus  raising  the  com- 
pression line. 

The  heavy,  full  line  curves  in  Fig.  19  give  steam  consumptions  plotted  against 
clearance  volumes,  for  different  mean  effective  pressures,  a  constant  length  of 


Fig.  21.  Saturated  steam  13  at.  abs.  Noncondensing. 
Most  favourable  compression. 


d      10 


Fig.  22.     Saturated  steam  13  at.   abs. 

Condensing  (0.1  at.  abs.). 
Most  favourable  compression. 


compression  of  90%,  saturated  steam  at  a  pressure  of  14  at.  abs.  and  atmospheric 
exhaust.  The  heavy  dashed  line  curves  also  give  the  theoretical  steam  consump- 
tions plotted  against  clearance  volumes  for  various  mean  effective  pressures,  but 
for  the  best  compression  in  each  case.  The  dashed-and-dotted  lines  are  lines  of 
constant  best  compression.  The  points  of  their  intersection  with  the  dashed  lines 
give  the  best  compression  for  that  particular  combination  of  M.E.P.  and  clearance 
volume. 

For  example,  an  M.E.P.  of  10  at.  requires  a  best  compression  of  20%  for 
a  clearance  of  12%,  the  steam  consumption  being  9.4  kg.    Also  for  6%  clearance, 


28 

an  M. E.P.  of  10  at.,  and  a  best  compression  of  10%,  the  steam  consumption  would 
be  8,8  kg. 

Fig.  20  shows  similar  curves  for  superheated  steam  of  14  at.  abs.,  a  tem- 
perature of  300°  C  and  atmospheric  exhaust. 

It  should  be  observed  that  in  Figs.  19  and  20  the  dashed  curves  for  an 
M. E.P.  of  2  at.  show  a  distinct  minimum.  At  this  point  the  expansion  reaches 
the  back  pressure.  Reduction  of  the  clearance  volume  beyond  the  point  in- 
dicated by  the  minimum  of  the  M. E.P.  curve,  for  a  constant  M.E.P.  of  2  at., 
results  in  a  loop  on  the  indicator  card  with  a  corresponding  increase  in  steam 
consumption. 


Fig.  23.      Superheated  steam  300°  C,  13  at.   abs. 

Noncondensing. 
Most  favourable  compression. 


Fig.  24.      Superheated  steam   300°  C, 

13  at.  abs.    Condensing  (0.1   at.   abs.) 

Most  favourable  compression. 


The  dashed  curves  of  Fig.  19  are  repeated  in  Fig.  21.  They  give  steam  con- 
sumptions for  different  mean  effective  pressures  plotted  against  clearance  volumes, 
for  saturated  steam  of  13  at.  abs.,  atmospheric  exhaust  and  best  compression  in 
each  case. 

Fig.  22  shows  similar  curves  for  a  back  pressure  of  0.1  at.  abs. 

Figs.  23  and  24  show  corresponding  curves  for  superheated  steam  of  14  at.  abs. 
and  a  temperature  of  300°  G,  Fig.  23  being  for  atmospheric  exhaust  and  Fig.  24 
for  condensing  operation  (back  pressure  0.1  at.  abs.). 

Apart  from  the  curves  for  low  mean  effective  pressures  it  is  interesting  to 
note  that  for  atmospheric  exhaust  the  steam  consumption  shows  an  almost  linear 


29 

dependence  upon  the  clearance,  the  best  compression  being  assumed  in  each  case. 
It  is  therefore  possible  to  calculate  the  mean  specific  volume  loss  per  1  %  clearance 
and  per  HP-hour.  For  dry  saturated  steam  of  14  at.  abs.,  this  is  found  to  be 
0.0918  kg/HP-hour  and  1%  clearance  for  atmospheric  exhaust,  and  0.1715  kg. 
HP-hour  and  1%  clearance,  for  condensing  operation.  The  corresponding  figures 
for  the  mean  specific  volume  loss  for  superheated  steam  of  14  at.  abs.  and  a  tem- 
perature of  300°  G  are  0.072  kg  for  atmospheric  exhaust,  and  0.12  kg  for  con- 
densing operation.  For  instance  in  the  case  of  a  single  cylinder  condensing  engine 
running  on  saturated  steam,  an  increase  in  clearance  volume  of  6%  will  raise  the 
steam  consumption  by  1  kg/HP-hour. 

The  curves  of  Figs.  21,  22,  23,  24  also  contain  the  steam  consumption  of  the 
ideal  engine  without  clearance,  which  is  given  by  the  intersection  of  the  M.  E.P. 
curves  with  the  zero  clearance  line.  The  distance  of  a  horizontal  line  drawn  through 
such  points  from  the  corresponding  M.  E.P.  curves  gives  the  volume  loss  for  any 
particular  clearance.  It  may  therefore  be  stated  that  for  the  same  initial  and  back 
pressure,  the  same  M.E.P.  and  best  compression,  the.  theoretical  steam  consumption 
and  the  volume  loss  increase  almost  linearly  with  the  clearance  volume.  Excluding 
small  clearances  and  mean  effective  pressures,  this  relation  is  strictly  true  for 
condensing  operation  and  approximately  so  for  atmospheric  exhaust. 

A  mathematical  expression  of  the  volume  loss  R  can  be  based  on  the  diffe- 
rence in  area  of  the  diagrams  with  and  without  clearance,  for  <p  =  constant  (see 
shaded  areas  in  Figs.  1,  2,  3);  ie,  pe,  iez  and  pe2  representing  the  total  heat  and 
pressure  at  the  end  of  expansion,  where  index  1  refers  to  the  engine  without,  and 
index  2  to  the  engine  with  clearance.  The  stroke  volume  is  again  assumed  to  be 
1  cbm.  Expansion  and  compression  are  adiabatic. 

1.  Arbitrary  length  of  compression  (counterflow  engines): 
R  =  427  (i,  —  ij  yi  <p  -f  10000  (Pet  —  pj  —  427  (i,  —  i  J  y,  (v  +  Vc)  - 

-  10000  (pet  -  p.)  (1  +  S0)  +  10000  pz  (!  —  &)  +  427  (ik  -  *2)  y2  (k  -f  S0)  + 

+  10000  fo-p.)  S0    ........     (VII) 

All  values  may  be  taken  from  the  Mollier  chart,  with  the  assumptions  that 
the  quality  of  the  steam  at  the  beginning  of  compression  is  equal  to  the  quality 
obtained  by  expanding  adiabatically  to  the  back  pressure,  and  that  the  specific 
volume  of  the  residual  steam  reduced  to  initial  pressure  is  equal  to  the  specific 
volume  obtained  by  continuing  the  compression  adiabatically  from  the  terminal 
compression  to  the  initial  pressure. 

For  an  approximate  calculation  with  almost  exact  results,  it  may  be  assumed 
that  iel  =  ie2  and  Pei  —  Pez,  or  approximately: 

R  =  10000  p2  (1  -  k)  +  427  (ik  -  i2)  •  y,  (A:  +  SJ  +  10000  (Pl  -  pk)  S0  - 


2.    100%  length  of  compression  (una-flow  engines): 
R  =  427  (i,  -  i,,)  Yl  -  <p  +  10000  (pei  —  p2)  —  427  (i,  —  iet)  n  (<p  +  V.)  - 

'  1-pfc)tS0     (IX) 


30 

Again,  for  approximate  results  it  may  be  assumed  that  iel  =  ie2  and  pel  —  pe2, 
or  approximately: 

R  =  427  (ik-i2)y2(i  +  S0)  +  10000  fo-p*)  S0  —  427  (*!  —  ij  ft  Vc- 

—  10000(pei  — p2)S0 (X) 

Vc  may  be  taken  directly  from  the  Mollier  chart.  The  result  R  of  this  cal- 
culation is  the  absolute  volume  loss  per  stroke  in  a  cylinder  having  a  clearance 
volume  of  S0,  a  stroke  of  1  m  and  an  area  of  1  sqm.  For  different  cylinder  sizes 
R  has  to  be  changed  proportionately.  The  HP  loss  may  be  obtained  by  multi- 
plying R  by  the  revolutions  per  second  and  dividing  by  75. 

D 

The  quotient   —   can  be  termed  "relative  volume  loss",  wherein 

£,,  =  427(1'!—  ie)  •  ft<p  + 10000  (pei  —  p2)     .....     (XI) 

and  is  the  output  per  stroke  in  mkg  of  a  cylinder  of  1  cbm  stroke  volume,  without 
clearance. 

For  a  cylinder,  with  clearance,  of  1  cbm  stroke  volume,  the  output  per  stroke 
in  mkg  would  be 

Lt  —  427  (ii  —  ij  ft  (<p  -\-  T7C)  -f- 10000  (pet  —  p2)  (1  +  S0)  —  10000  pz  (l  —  k)  — 

_427(4  —  i2)yz  (k  +  S0)  — 10000(^  —  ^)^0  .     .     .     .     (XII) 
This  calculation  gives  the  relative  volume  loss  for  una-flow  engines  having 
100%  compression  and  95%  vacuum,  referred  to  the  output  per  stroke  of  an  •  » 
ideal  engine  without  clearance.    This  relative  volume  loss  is  almost  independent 
of  the  initial  pressure. 

For  initial  pressures  of  8  to  15  at.  abs.,  this  gives 

for  a  clearance  of  1%,  a  mean  relative  volume  loss  of    5% 

5555  55  55        On/          '5  55  55  55  55  55  CO/ 

«  /o»  °  /o 

55         55  55  55       Q  O/          "  "  "  "  "  "  ft  I/   O/ 

•3    /Ol  °   /2  /O 

55          55  55  55         /   O/  "  "  "  "  "  "        4   4    I/O/ 

*7o>  ii/2/o 

55??  55  "KO/  "  "  "  "  "  "        1  A  °/ 

3    /O5  **  /O 

?5          ??  55  55        HO/  5?  55  55  55  55  55        O  \    Q/ 

'     /05  Zi    /O 

55         55  55  55       Q  O/  "  "  "  "  "  "       07  O/ 

These  figures  also  show  an  approximately  linear  relation  between  volume 
loss  and  clearance  volume,  except  for  very  small  clearances. 

As  a  final  result  of  the  foregoing  discussions,  Fig.  25  shows  a  simple  rule 
which  allows  the  best  compression  to  be  determined  for  any  given  case.  For 
a  given  amount  of  steam  (p  the  compression  must  evidently  be  correct  if  a  displace- 
ment of  the  line  (p  by  an  amount  dq>  produces  equal  changes  of  area,  shown  shaded 
in  Fig.  25,  on  both  the  expansion  and  compression  sides  of  the  diagram,  in  such 
a  way  that  the  following  equation  is  satisfied: 

427  (ij  —  ie)  •  d<p  •  7X  =  427  (ih  —  i*2)  d<p  'ft     or:    i±  —  ie  =  ih  —  i2  •     •      •   (XIII) 

or  in  other  words  the  change  of  total  heat  during  expansion  is  equal  to  the  change 
of  total  heat  during  compression.  It  is  therefore  only  necessary  to  add  the  change 
of  total  heat  during  expansion  to  the  total  heat  corresponding  to  the  back  pres- 
sure, in  order  to  obtain  the  best  terminal  compression  pressure.  By  laying  out  the 
corresponding  compression  line,  the  best  length  of  compression  will  be  determined. 


31 


f— ^ 


It  again  follows  that  for  100%  cut-off  the  compression  is  zero,  and  for  ex- 
pansion to  the  back  pressure  the  compression  must  reach  the  initial  pressure. 

The  following  fundamental  law  therefore  applies:  For  given  initial  pressure, 
back  pressure,  mean  effective  pressure  and  clearance  volume,  the  length  of  compression 
must  be  such  that  the  change  of  total  heat  during  expansion  is  equal  to  the  change  of 
total  heat  during  compression. 

Interchanging  the  words  "back  pressure"  and  "length  of  compression"  with 
each  other  again  leads  to  equality  of  heat  change.  In  other  words,  for  a  given 
initial  pressure,  mean  effective  pressure, 
clearance  volume  and  length  of  compression, 
the  back  pressure  must  be  such  that  the 
change  of  total  heat  during  expansion  is 
equal  to  the  change  of  total  heat  during 
compression. 

Starting  with  a  diagram  of  equal  heat 
changes,  a  reduction  in  back  pressure, 
with  the  same  clearance  and  the  same 
length  of  compression,  produces  a  smaller 
heat  change  during  compression  (seeMollier 
chart),  and  for  the  same  initial  pressure 
and  M.E.P.,  a  larger  heat  change  during 
expansion,  with  the  result  of  an  increase 
in  steam  consumption.  An  increase  in 
back  pressure  under  the  same  assumptions 
produces  inverse  heat  changes  and  also 


Fig.  25. 


increases  the  steam  consumption.  The  inherent  change  in  steam  consumption 
caused  by  a  reduction  in  back  pressure  is  an  entirely  separate  consideration. 

An  interchange  of  the  words  "initial  pressure"  and  "back  pressure"  again 
leads  to  equality  of  total  heat  changes,  or  in  other  words,  for  a  given  back  pressure, 
mean  effective  pressure,  clearance  volume  and  length  of  compression,  the  initial  pres- 
sure must  be  such  that  the  change  of  total  heat  during  expansion  is  equal  to  the  change 
of  total  heat  during  compression. 

Starting  again  with  a  diagram  of  equal  heat  changes,  a  reduction  in  initial 
pressure  also  reduces  the  change  of  total  heat  during  expansion  for  the  same  length 
of  compression  and  the  same  back  and  mean  effective  pressures.  Inversely,  an 
increase  in  initial  pressure  will  also  increase  the  heat  change  during  expansion. 
The  heat  change  during  compression  still  being  the  same,  an  increase  in  steam 
consumption  will  result  on  account  of  the  inequality  of  the  heat  changes.  The 
variation  of  steam  consumption  due  to  a  change  in  initial  pressure  alone  is  a  point 
to  be  considered  separately. 

The  wording  of  the  last  two  forms  of  this  fundamental  law  has  particular 
reference  to  compound  and  triple  or  quadruple-expansion  engines.  For  instance, 
a  variation  of  cut-off  in  the  low  pressure  cylinder  of  a  compound  engine  affects 
the  exhaust  pressure  of  the  high  pressure  cylinder  as  well  as  the  admission  pres- 
sure of  the  low  pressure  cylinder,  and  renders  possible  an  even  distribution  of 
heat  changes  in  both  cylinders. 


32 


Considering  the  high  and  low  pressure  diagrams  as  one,  this  leads  to  equal 
changes  of  total  heat  during  the  combined  expansion  and  compression,  i.  e.  equal 
heat  changes  in  each  part  diagram  result  in  equal  heat  changes  in  the  combined 
diagram. 

The  same  reasoning  applies  to  triple  and  quadruple-expansion  engines. 
A  different  problem  is  presented  by  the  determination  of  the  best  clearance 
volume  for  a  given  M.E.P.  and  given  length  of  compression.    This  would  apply 
especially  to  una-flow  engines  where  the  length  of  compression  is  fixed. 

The  heavy  lines  in  Figs.  19  aftd  20  correspond  to  constant  compression  (90%), 
and  intersect  the  dashed  lines,  these  intersections  being  located  on  the  dashed- 
and-dotted  line  corresponding  to  a  best  compression  of  90%.  Each  intersection 

refers  to  a  definite  clearance  volume  and 
mean  effective  pressure  for  which  com- 
bination 90%  compression  is  the  best.  The 
heavy  lines  indicate  that  with  a  constant 
compression  of  90%  the  steam  consumption 
can  be  diminished  by  reducing  the  clearance. 
Referring  to  Fig.  20,  and  taking  for  in- 
stance, the  dashed  curve  for  pt  =  2  at. 
representing  steam  consumption  with  best 
compression,  it  will  be  found  that  for  17,% 
clearance  the  best  compression  is  90%. 
Following  the  full  line  curve  for  pt  =  2  at., 
passing  through  this  point,  a  clearance  vo- 
lume of  10%  is  found  to  give  the  lowest 
steam  consumption  with  90%  compression 
and  this  mean  effective  pressure.  It  is  ob- 

vious that  at  this  point  the  heat  changes  during  expansion  and  compression  are 
not  equal,  since  this  occurred  for  a  clearance  volume  of  17%. 

The  rule  for  determining  the  best  clearance  volume  for  a  given  M.E.P.  and 
length  of  compression  can  be  arrived  at  by  the  following  reasoning. 

The  full  line  diagram  in  Fig.  26  is  assumed  to  have  equal  heat  changes  during 
expansion  and  compression.  It  is  then  possible  to  increase  the  area  of  the  dia- 
gram by  transposing  the  lines  bounding  the  stroke  volume  towards  the  left,  the 
points  of  cut-off  and  compression  remaining  fixed.  On  account  of  this  transposition, 
the  new  diagram  has  a  slightly  longer  compression.  A  reduction  of  the  length  of 
compression  in  the  new  diagram  to  the  original  value  will  diminish  the  pressure 
change  during  compression,  and,  for  the  same  amount  of  steam  admitted  also 
increase  the  pressure  change  during  expansion;  the  diagram  area  still  remaining 
increased.  Repeating  this  procedure,  the  point  of  best  clearance  volume  is  ap- 
proached, for  which  the  changes  of  total  heat  during  compression  and  expansion 
are  equal.  This  theorem  has  been  proved  here  for  a  given  constant  quantity  of 
steam  admitted.  Since  it  is  the  same  problem  whether  the  maximum  M.E.P. 
is  required  for  a  given  quantity  of  steam  admitted,  or  the  maximum  quantity  of 
steam  for  a  given  M.E.P.,  the  rule  may  be  stated  thus:  for  given  initial  pres- 
sure, back  pressure,  mean  effective  pressure  and  length  of  compression,  the  clearance 


Fig.  26. 


33 


volume  must  be  such  that  the  pressure  change  during  expansion  is  equal  to  the  pressure 
change  during  compression.  Under  certain  conditions,  therefore,  an  increase  in 
clearance  volume  may  cause  a  reduction  in  steam  consumption.  Nearly  all  una- 
flow  engines  violate  this  rule.  The  majority  of  condensing  engines  have  too  large, 
and  non-condensing  engines  too  small  a  clearance  volume.  Referring  to  Fig.  27, 
and  starting  as  before  with  equality  of  changes  of  total  heat,  then  for  the  same 
terminal  expansion  and  compression  pressure  and  different  clearance  volumes  it 
follows  that,  approximately, 
*i  sz  VK 


j  — |- 


or 


-f  •?!  &2  = 


c  A* 

or   ~~  =  ~ 


This  means  that  for  the  same  initial  and  back  pressure,  the  same  terminal  expan- 
sion and  compression  pressure,  and  equality  of  heat  changes  during  expansion  and 
compression,  the  lengths  of  the  best  com- 
pressions have  the  same  ratio  as  the 
clearance  volumes. 


III 


Fig.  27. 


Fig.  28. 


A  withdrawal  of  steam  during  expansion  necessarily  changes  the  length  of 
compression.  As  will  be  shown  later,  una-flow  engines  permit  the  bleeding  of 
steam  during  expansion.  One  or  two  withdrawals  according  to  Fig.  28  will  shift 
the  beginning  of  compression  from  I'  to  II'  or  IIP.  The  same  effect  will  be  pro- 
duced by  an  increase  in  exhaust  lead. 

Finally,  a  comparison  of  Figs.  29  and  30  shows  the  favorable  manner  in 
which  compounding  influences  the  volume  loss.  The  diagram  of  Fig.  29  has  the 
same  clearance  volume  as  the  low  pressure  cylinder  in  Fig.  30.  The  comparison 
reveals  the  fact  that  the  volume  loss  of  the  low  pressure  part  of  both  diagrams 
is  about  the  same,  but  that  for  the  high  pressure  part  the  compound  diagram 
shows  a  decidedly  smaller  volume  loss.  It  should  not  be  overlooked,  however, 
that  the  una-flow  diagram  can  be  realized  in  a  cylinder  with  only  1%  clearance 
(single  beat  valves),  while  low  pressure  cylinders  of  ordinary  design  have  a  clea- 
rance of  about  6%  or  more. 

Fig.  30  further  indicates  that  the  greatest  part  of  the  volume  loss  in  com- 
pound engines  occurs  in  the  low  pressure  cylinder,  and  comparatively  little  in 
the  high  pressure  cylinder. 

Cylinder  jacketing  in  una-flow  engines  causes  a  considerably  steeper  com- 
pression line  as  well  as  higher  expansion  line,  which  must  be  compensated  for 

Stumpf,  The  una-flow  steam  engine.  3 


34 


-Ae 


by  shortening  the  length  of  compression.  Leaky  inlet  valves 
will  have  the  same  effect.  Leaky  exhaust  valves,  however, 
would  require  longer  compression  to  compensate  for  the  re- 
sultant change  in  the  expansion  and  compression  lines. 

The  maximum  volume    loss  occurs    when   the  admission 
is  longer  than  the  compression,    and    part 
of    the    clearance    volume    is   arranged    on 
the    cylinder    at    a   point  between    end    of 
admission    and    beginning    of   compression, 
because   under  these   conditions    the    com- 
pression    is    unable     to 
correct    the    losses    due 
to  this  clearance. 

Decrease  of  this  clea- 
rance volume  loss  occurs, 
if  the  clearance  space  is 

placed  nearer  to  the  inlet 
Fi°-.  29. 

end   of  the    cylinder    as 

regards  the  compression  part  of  the  stroke  but  nearer  to  the  exhaust 
\        end   of  the  cylinder  as  regards  the  expansion  part  of  the  stroke. 

Critical  Back  Pressure. 

The  shape  of  the  steam  consumption  curves  in  Fig.  18 
indicates  that  under  certain   conditions   a  length  of  com- 
pression  in  excess   of  100%   is  the  most  favorable.      In   a 
steam  cylinder,  however,   only  compressions  up   to   appro- 
ximately 100%  can  be  realized.    A  best  compression 
of  120%  therefore  would  mean  a  rise  in  back  pres- 
sure  corresponding  to    20%    compression    previous 
to  the  commencement  of  the  cylinder  compression 
of    100%.    This   increase    in    back    pressure    could 

be  obtained  by  some 
kind  of  throttling 
organ  in  the  exhaust 
pipe  or  in  the  cool- 
ing water  pipe  or 
by  admitting  air 
into  the  condenser. 

In  other  words,  if  this  higher  back  pressure  is  not  in  effect,  the  compression 
remains  at  100%  with  a  resultant  increase  in  steam  consumption  due  to  wrong 
length  of  compression.  If  now  the  back  pressure  is  raised  an  amount  correspon- 
ding to  20%  compression,  this  back  pressure  would  be  the  best  for  100%  com- 
pression. A  further  increase  of  the  back  pressure  beyond  this  point  would  result 
in  an  increased  steam  consumption  for  100%  compression.  This  leads  to  the 
conception  of  the  therm  "critical  back  pressure".  The  critical  back  pressure, 
therefore,  is  that  value  of  the  back  pressure  at  the  beginning  of  compression,  the 


Fig.  30. 


increase  or  decrease  of  which  for  the  same  length  of  compression,  the  same 
clearance  volume,  the  same  mean  effective  pressure  and  the  same  initial  pressure, 
results  in  an  increased  steam  consumption.  Fig.  18  shows  the  effect  of  the 
compression,  and  Fig.  31  the  effect  of  the  output  (mean  effective  pressure)  upon 
the  critical  back  pressure. 

Fig.  31  gives  the  theore- 
tical steam  consumption  for 
90  %  compression  plotted 
against  the  back  pressure.  All 
three  curves  slope  towards 
the  right,  i.  e.,  the  steam  con- 
sumption diminishes  with  in- 
creasing back  pressure.  Thrott- 
ling of  the  exhaust  or  partial 
reduction  of  the  vacuum  would 
thus  reduce  the  steam  con- 
sumption. The  critical  back 
pressure,  therefore,  has  been 
exceeded  in  all  three  cases.  At 
the  same  time  the  influence 
of  the  output  can  be  quite 
clearly  seen,  since  the  slope 
of  the  curves  for  early  cut- 
offs is  more  pronounced  than 
it  is  for  the  others.  The  cri-  Fig.  31.  Superheated  steam  300°  C,  13  at.  abs. 
tical  back  pressures  will  be  Clearance  space  5%. 

reached    when    these    curves 

attain  their  lowest  points  and  reverse  their  slope.  The  uppermost  curve  will  reach 
its  minimum  sooner  than  the  lower  curves ;  or  in  other  words,  the  critical  back 
pressure  is  the  higher,  the  lower  the  output  (MEP). 

The  starting  point  for  the  calculation  of  the  critical  back  pressure,  according 
to  the  above,  is  not  the  cut-off  but  rather  the  output  or  MEP.;  or,  for  given 
initial  pressure,  clearance  volume  and  length  of  compression,  the  amount  of  steam 
admitted  is  the  basis.  According  to  Fig.  5,  and  referring  to  page  18,  the  reciprocal 
of  the  value  of  the  steam  consumption  for  adiabatic  expansion  and  compression 

13  =  const.)  is: 


0,os 
Abso/ufer  0rvc/r  /fn 

Condenser  fires-sure 


0, 10 


E  — 


0  = 


E2f^  (tV--  F*'-«)-*LL^-(TV 


A-n\_P2VKn(y    i_B_ 


9 


PI     VE    L     !vE\n- 

<p  '  n  —  1  [         \VH] 


Pi     <P     n  — 


3* 


36 


=±V     -!<:  4-1  -i-  2    Pi  VK  _ 

<p  <p  y'  n  —  1        <p  (rc-i)T-V1-1    "  Pi'  (p  (n  —  i)~S0n~l  ~ 

I    Pa    Pi    VK         p2   Piv     |    p2    Pi   v 

t    ~  '  '  H     |  '         '   '  K- 

P!    <p  n  —  i        P!    <p  p1    (p 

~ 
n 


Now     VE  =  cp-\-  Vc  =  <p  -\-  VK  \  —  )n  ;    Substituting   this    in   the  equation  for  <9, 


_P^Pi_       VK  .    pz    Pi    ^g-^        P2PiT/ 

Pi  9?  (w—  l)^"1        Pi'  9>  '  w-1  "      Pi  ^ 

The  bracketed  part  of  the  fifth  term  can  be  expressed  as  a  series  by  the  bino- 
mial theorem;  considering  the  first  three  terms  only,  the  expansion  gives: 


Substituting  this  value  in  the  equation  for  <9, 


«,     o  ~  „       r       „       /IT/ 
OP  n  —  1        ft  —  *-\V  HI 

\  / 

^/  ^  r-1  /P»W»  .  /^\2/  <P  Y1-1 


Pl/       ^    '      \    9    /    \^H, 

T  H. 


If  the  steam  consumption  is  to  be  a  minimum  or  0a  maximum,  it  follows  that 
d& 


-  —  0. 

r/£l\ 

HpJ 

Then 

d^         /P2\i^   Pi    Fjr         /p2\L^      Pl  F^     /  cp 


Pi  <p  n  —  1      \p1/  (n  —  i)<p\VH 


/Fx\«  /j^X-^Pi  ^^  pi  /^ 

l  V  j    i^H/  V  (n-l)V-1       V         ^-1 


_i         _ 


or 


l-n  2-n 


....  (XV) 

If  VH=  VK  i.  e.  for  100%  length  of  compression  (una-flow  engines),  this  for- 
mula simplifies  to: 

l-n 
/PA     » 
\Pl/ 


37 

1— n 

Substituting  in   these  equations    [  — J  -  and   n  =  1,33    or   -^-,  then 

n-i  W 


or  x.' 

Pi]  \Pi) 

2  — n  1  1  2  — n  2  —  1,33 

rr~ii  n  /     r   2  1          Tl  1  /     r   2  \  71  1  f\   \  ft   1  \      3  ^    —   \  O 

I    nOT'OtnT'O       I  — -  T* —  -*-    •    J  *!*""—•     •!*  — '—    *!*  -1    — —     /v»   A>°  «*  *     /y»6 

lilt/lclUlt;      I  »*/  **/  »**  JU  JL/  — —  ,//    w 

\ft/  \^v 

Hence    equation    (XV)    can   be   expressed   in   the  form:      ax2  —  c  = 

x 

or     x3  -I a; =  0. 

a  a 

This  cubic  equation  may  be  written  x3-}-px-{-q  =  Q  where  p  =  —  and 

Cv 

=  —  — ,  and  may  be  solved  by  Cardani's  formula  as  follows: 


\ 


With  the  help  of  equations  XV,  XVI  and  XVII,  the  critical  back  pressures 
p2  have  been  figured  for  various  initial  pressures,  clearance  volumes,  lengths  of 
compression,  and  amounts  of  admitted  steam.  The  results  are  plotted  in  Figs.  32, 
33,  34  and  35,  and  show  the  influence  of  these  different  variables  upon  the  critical 
back  pressure. 

Fig.  32  gives  the  variation  of  the  critical  back  pressure  for  different  clearance 
volumes  and  initial  pressures,  for  a  constant  amount  of  steam  admitted  (p  =  10% 
and  a  constant  length  of  compression  of  100%  (una-flow  steam  engine).  The  ordi- 
nates  of  the  individual  curves  indicate  that  for  a  given  clearance  volume  the  cri- 
tical back  pressure  pz  is  proportional  to  the  initial  pressure  p^  which  is  also  evi- 
dent from  the  previous  equation, 

1__  1  33 


=-,    or   pz  =  p-n=pix--*=pl*   .     (XVIII) 


Therefore,  for  the  same  clearance  volume,  the  same  output  and  the  same  length 
of  compression,  the  critical  back  pressure  is  proportional  to  the  initial  pressure. 

For  the  same  initial  pressure,  output  and  length  of  compression,  the  critical  back 
pressure  varies  with  the  clearance  volume  at  a  steadily  increasing  rate,  at  first  slowly, 
then  more  rapidly,  until  the  rate  of  increase  attains  a  linear  maximum. 

The  critical  back  pressure  is  zero  for  zero  clearance  volume,  and  is  small  for 
very  small  clearances.  This  is  self-evident  and  also  proved  by  the  curves,  but 
is  frequently  left  out  of  consideration  in  the  design  of  steam  engines.  The  clea- 
rance volume  is  therefore  the  cause  of  all  evil.  If  the  clearance  volume  is  made 
zero  the  critical  back  pressure  as  well  as  the  volume  loss  are  zero,  and  the  reci- 
procating engine  in  this  respect  is  put  on  the  same  level  as  the  steam  turbine. 
Noteworthy  is  the  slow  initial  increase  of  the  critical  back  pressure,  which 
again  calls  attention  to  the  necessity  of  small  clearance  volumes. 


38 


Fig.  33  gives  the  relation  of  the  critical  back  pressure  to  the  amount  of  steam 
admitted,  for  different  clearance  volumes,  for  a  constant  initial  pressure  of  13  at.  abs. 
and  a  constant  length  of  compression  of  100%.  The  critical  back  pressure  grows 
with  decreasing  amount  of  steam  admitted.  In  this  case  also  the  necessity  for 
small  clearance  volumes  is  evident,  especially  for  early  cut-offs,  as  for  instance 
in  condensing  una-flow  engines. 

Therefore:  for  a  constant  clearance  volume,  for  the  same  initial  pressure  and 
length  of  compression,  the  critical  back  pressure  increases  with  decreasing  output. 


Fig.  32. 

Critical  back  pressure  plotted  against  length  of  compression  is  shown  in  Fig.  34. 
Each  curve  refers  to  a  different  amount  of  steam  admitted,  the  initial  pressure 
being  13  at.  abs.  and  the  clearance  volume  2%  in  all  cases.  The  curves  show  first 
a  rapid  increase  of  the  critical  back  pressure  which  attains  a  maximum  at  about 
30  or  40%  length  of  compression,  followed  by  a  steady  decrease.  The  effect  of 
the  compression  is  the  less,  the  higher  the  MEP  or  the  greater  the  amount  of 
steam  admitted.  The  curves  confirm  the  lengths  of  compression  ordinarily  used 
in  counterflow  engines  for  normal  cut-offs  of  30  to  40%  and  more,  and  the  com- 
pression of  una-flow  engines  (100%)  for  the  usual  admissions  of  15  to  10%  and 
less.  On  the  other  hand  Figs.  32  and  33  seem  to  lend  support  to  the  long  admis- 
sions and  subdivided  pressure  ranges  of  counterflow  engines.  In  the  case  of  multi- 
stage engines  the  low  pressure  cylinder  and  receiver  pressure  are  the  determining 
factors  for  the  critical  back  pressure.  The  compression  has  the  least  influence, 
especially  for  late  cut-offs.  It  should  be  noted  here  that  the  scale  of  ordinates  in 
Fig.  34  is  twice  that  of  Figs.  32,  33  and  35. 


39 


The  interrelation  of  critical  back  pressure,  length  of  compression  and  clea- 
rance volume  is  given  in  Fig.  35,  the  initial  pressure  being  13  at.  abs.  and  the 
amount  of  steam  admitted  10%.  The  curves  show  the  immense  influence  of  the 
clearance  upon  the  critical  back  pressure,  which  latter  seems  to  follow  a  geometrical 
progression  with  increasing  clearance.  The  critical  back  pressure  increases  rapidly 
up  to  about  40%  length  of  compression,  the  rate  of  increase  being  progressively 
larger  for  larger  clearances,  after  which  it 
shows  a  steady  decrease  with  increasing  length 
of  compression. 

In  the  order  of  the  influence  exerted,  the 
clearance  volume  ranks  first,  then  initial  pres- 
sure, then  admission  and  finally  length  of  com- 
pression. The  need  for  small  clearance  volume 


9.0 


too 


Fig.   34. 


loo 


is  imperative.  This  is  fulfilled  in  the  best  pos- 
sible manner  by  the  una-flow  engine,  especially 
if  fitted  with  high  lift  single-beat  poppet  valves 
which  allow  a  clearance  of  1%  to  be  realized. 
The  length  of  compression  (90%)  combined 
with  the  short  cut-offs  as  used  in  una-flow 
engines  is  also  favorable,  since  Fig.  32  gives 
a  critical  back  pressure  of  only  0.004  at.  abs., 
for  P!  =  10  at.  abs.,  9  =  10%  7*=  100%, 

and  SQ  =  1  % ;  a  back  pressure  which  is  beyond  even  the  most  modern  condensing 
equipment.  At  the  same  time  the  above  combination  also  gives  a  very  small 
volume  loss.  The  question  is  not  to  produce  an  engine  which  has  the  smallest 
critical  back  pressure,  but  one  which  combines  the  latter  with  a  small  volume 
loss.  It  is  possible  to  have  a  large  volume  loss  and  yet  a  critical  back  pressure 
equal  to  zero,  for  instance  in  the  case  of  large  clearance  volume  and  compres- 
sion equal  to  zero. 


40 

On  account  of  the  adverse  influence  of  high  initial  pressure  combined  with 
short  admission,  the  single  stage  una-flow  engine  has  to  rely  upon  small  clearance 
volumes  which,  however,  can  be  realized  without  difficulty.  Compound  or  triple 
expansion  counterflow  engines  can  be  run  with  more  liberal  clearance  volumes 
by  reason  of  the  long  admission  and  the  low  initial  or  receiver  pressure  and  still 
have  a  critical  back  pressure  beyond  that  attainable  with  the  best  condensers. 
The  case  is  very  different  for  single  stage  counterflow  engines  which  usually  have 
very  large  clearance  volumes.  For  instance,  according  to  Fig.  35,  a  single  stage 
engine  with  iS"0=10%,  pl  =  13  at.  abs.,  99  =  10%,  and  Ffc  =  40%,  has  a  cri- 
tical back  pressure  of  about  0.5  at.  abs.  All  the  bad  influences  are  cumulative; 
large  clearance  volume,  short  admission  and  the  most  unfavorable  length  of  com- 
pression of  40%.  It  might  not  be  impossible  to  find  an  engine  combining  all  these 
points,  in  actual  operation.  For  20%  clearance,  15  at.  abs.  initial  pressure,  10% 
admission  and  40%  length  of  compression,  the  critical  back  pressure  becomes 
1  at.  abs.  To  run  such  an  engine  condensing  would  be  utterly  wasteful.  The  use 
of  several  stages  not  only  reduces  the  volume  loss,  but  also  lowers  the  critical  back 
pressure,  and  to  such  a  degree  that  other  defects,  such  as  large  clearance  volumes,  lose 
their  significance,  to  a  certain  extent. 

As  can  be  seen  from  Figs.  32,  33,  34  and  35,  the  critical  back  pressure  becomes 
zero  for  S0  =  0,  ^  =  0,  Vk  =  0  and  attains  a  maximum  for  zero  admission.  In 
all  the  above  considerations,  admission,  mean  effective  pressure  and  output  have 
the  same  meaning  and  may  be  used  indiscriminately. 

The  possibility  of  causing  an  increase  in  steam  consumption  by  going  beyond 
the  critical  back  pressure,  as  well  as  the  useless  generation  of  too  high  a  vacuum 
are  out  of  the  question  in  case  of  well  designed  una-flow  engines.  These  condi- 
tions, however,  sometimes  occur  in  counterflow  engines,  even  to  such  an  extent 
that  the  engineer  and  fireman  are  able  to  notice  the  bad  effect  of  too  high  a  vacuum. 
Prof.  Josse  reports  such  a  case  in  the  Zeitschrift  des  Vereines  deutscher  Ingenieure, 
1909,  page  324.  He  states  that  "the  economy  of  the  engine  improved  until  the 
back  pressure  fell  to  0.2  at.  abs.  From  this  point  onwards  a  further  reduction 
in  steam  consumption  due  to  increased  vacuum  was  not  noticeable".  Such  a  result 
in  this  particular  engine  was  caused  not  only  by  the  critical  pressure  being  exceeded, 
but  also  by  increased  losses  of  initial  condensation  and  leakage  due  to  the  higher 
vacuum.  The  initial  condensation  was  considerable  in  this  case.  Further  more, 
the  pressure  difference  between  the  engine  cylinder  and  condenser  increases  with 
a  high  vacuum  by  reason  of  the  usual  deficiency  in  exhaust  area.  In  una-flow 
engines  the  exhaust  port  areas  can  always  be  made  sufficiently  large;  leakage  and 
initial  condensation  are  reduced  because  the  engine  has  no  exhaust  valves  and 
may  be  fitted  with  single-beat  inlet  valves,  and  furthermore  has  the  benefit  of 
the  una-flow  action.  In  well  designed  and  well  built  una-flow  engines  the  results 
of  the  above  calculations,  which  implicitly  contain  the  rule  of  equal  heat  changes, 
do  not,  therefore,  require  any  appreciable  corrections. 

For  given  initial  and  mean  effective  pressures,  the  lowest  steam  consumption 
is  obtained  when  the  clearance  volume  is  zero  and  the  back  pressure  is  zero;  the 
length  of  compression  has  then  no  influence.  The  critical  back  pressure,  i.  e.  the 
most  economical  back  pressure,  and  the  length  of  compression  become  of  impor- 


41 

tance  as  soon  as  the  clearance  volume  has  a  definite  value.  According  to  Fig.  31 
the  length  of  compression  can  then  be  increased  to  100%  and  the  back  pressure 
reduced  until  the  changes  of  total  heat  during  compression  and  expansion  become 
equal,  thus  offering  the  best  basis  for  low  steam  consumption.  In  other  words: 
for  a  given  initial  pressure,  given  mean  effective  pressure  and  given  clearance  volume, 
the  minimum  steam  consumption  will  be  obtained  when  the  length  of  compression  is 
100%  and  the  back  pressure  is  such  that  the  change  of  total  heat  during  expansion 
is  equal  to  the  change  of  total  heat  during  compression.  (Una-flow  steam  engine.) 

This  rule  may  also  be  arrived  at  if  the  second  wording  of  the  fundamental  law 
given  on  page  31,  dealing  with  back  pressure,  is  applied  to  una-flow  engines.  The 
minimum  steam  consumption  requires  the  shortest  possible  cut-off  and  the  longest 
possible  compression  of  100%,  these  being  related  to  each  other  by  the  rule  of 
equal  heat  changes. 

An  examination  of  the  common  types  of  steam  engines  will  reveal  the  fact 
that  incorrectly  designed  engines  are  the  rule  and  correctly  designed  engines  the  ex- 
ception. There  is  hardly  a  steam  engine  designer  who  is  not  guilty  of  some  viola- 
tion in  this  respect.  To  begin  with,  the  average  designer  is  not  aware  of  the  harm- 
fulness  of  the  clearance  volume,  which  explains  the  carelessness  with  which  un- 
necessarily large  clearances  are  used.  The  latter  are  rendered  necessary  for  short 
cut-offs,  for  instance  in  locomotives,  marine  engines,  or  non-condensing  engines 
with  shaft  governor  controlling  both  inlet  and  exhaust.  In  marine  engines  this 
is  the  case  with  the  high  and  intermediate  cylinders,  while  the  low  pressure  cylin- 
ders usually  have  unnecessarily  large  clearances.  The  lengths  of  compression  are 
frequently  incorrect.  These  "necessarily"  and  "unnecessarily"  large  clearances 
can  be  avoided.  A  knowledge  of  the  above  rules  is  indispensable,  as  well  as  re- 
cognition of  the  fact  that  changes  in  load  as  well  as  change  of  rotation  may  be 
accomplished  merely  by  the  steam  admission  organs  without  change  in  the  exhaust 
timing.  The  proper  choice  of  steam  distributing  organs  as  well  as  their  arrange- 
ment and  mechanism  are  also  important  factors.  For  instance,  a  single-stage 
condensing  una-flow  marine  engine  with  single-beat  valves  arranged  in  the  cylinder 
heads  fulfills  all  of  the  above  conditions.  (See  Fig.  31,  Chapter  V.)  This  engine 
has  the  great  advantage  of  a  small  clearance  volume  of  less  than  1  % ;  the  exhaust 
timing  is  independent  of  the  inlet  gear  and  the  constant  length  of  compression 
is  correct  and  permissible.  The  same  reasoning  holds  true  for  the  stationary  una- 
flow  engine  having  single-beat  valves.  (See  Fig.  6,  Chapter  VI.)  Both  types  of 
engine  therefore  have  very  small  volume  losses  despite  the  large  pressure  ranges. 
In  the  same  way  a  considerable  reduction  of  clearance  volume  in  non-condensing 
una-flow  engines  can  be  accomplished  by  shortening  the  compression  (large  exhaust 
lead).  This  is  shown  in  Fig.  32,  Chapter  III,  including  also  the  effect  of  an  exhaust 
ejector,  which  produces  a  proper  change  of  compression  with  the  cut-off,  thereby 
further  reducing  the  volume  losses. 

Summary. 

1.  The  volume  loss  is  determined  by  the  clearance  volume,  initial  pressure, 
back  pressure,  mean  effective  pressure  and  length  of  compression.  It  increases 
with  increasing  initial  pressure  and  clearance  volume,  decreases  with  increasing 


42 

back  pressure  and  mean  effective  pressure,  and  becomes  a  minimum  for  a  certain 
length  of  compression. 

2.  Correct  compression  tends  to  reduce  the  volume  loss;  compression  may 
be  kept  constant  for  small  clearance  volumes,  but  should  be  varied  inversely  with 
the  cut-off  in  case  of  large  clearances  (Single  cylinder  engines). 

3.  Change  of  compression  ,in  case  of  high  vacuum  has  no  material  effect  upon 
the  steam  consumption. 

4.  The  clearance  volume  loss  is  zero  if  expansion  reaches  the  back  pressure 
and  compression  rises  to  the  initial  pressure. 

5.  The  theoretical  steam  consumption,   for  the  same  initial  pressure,  back 
pressure,  mean  effective  pressure  and  best  compression  in  each  case,  increases 
nearly  linearly  with  the  clearance  volume.    Apart  from  very  small  values  of  the 
clearance  and  mean  effective  pressure,  this  linear  dependence  is  almost  exact  for 
condensing  operation,  and  approximate  for  atmospheric  exhaust. 

6.  For  given  initial  pressure,  back  pressure,  mean  effective  pressure  and  clea- 
rance volume,  the  length  of  compression  must  be  such  that  the  change  of  total 
heat  during  expansion  is  equal  to  the  change  of  total  heat  during  compression. 

7.  For  given  initial  pressure,  mean  effective  pressure,  clearance  volume  and 
length  of  compression,  the  back  pressure  must  be  such  that  the  change  of  total 
heat  during  expansion  is  equal  to  the  change  of  total  heat  during  compression. 

8.  For  given  back  pressure,  mean  effective  pressure,  clearance  volume  and 
length  of  compression,  the  initial  pressure  must  be  such  that  the  change  of  total 
heat  during  expansion  is  equal  to  the  change  of  total  heat  during  compression. 

9.  For  given  initial  pressure,  back  pressure,  mean  effective  pressure  and  length 
of  compression,  the  clearance  volume  must  be  such  that  the  change  of  pressure 
during  expansion  is  equal  to  the  change  of  pressure  during  compression. 

10.  For  the  same  initial  pressure,  back  pressure,  the  same  terminal  compres- 
sion pressure  and  terminal  expansion  pressure,   and   for  equality  of  total  heat 
changes,  the  lengths  of  the  best  compressions  have  the  same  ratio  as  the  clearance 
volumes.    (Different  mean  effective  pressures.) 

11.  With  proper  proportioning  of  the  length  of  compression,  the  clearance 
volume  has  to  be  kept  as  small  as  possible;  this  applies  especially  to  single  cylinder 
condensing  engines  and  the  low  pressure  cylinders  of  compound  and  triple  expan- 
sion engines. 

12.  Subdivision  into  stages  results  in  reduction  of  volume  losses,  the  high 
pressure  cylinder  having  the  smallest  and  the  low  pressure  cylinder  the  largest 
volume  loss.    Intermediate  cylinders  have  a  loss  between  both  according  to  their 
relative  size. 

13.  For  given  initial  pressure,  mean  effective  pressure  and  clearance  volume, 
the  lowest  steam  consumption  is  obtained  if  the  length  of  compression  is  made 
100%  and  the  back  pressure  chosen  so  as  to  make  the  change  of  total  heat  during 
expansion  equal  to  the  change  of  total  heat  during  compression  (una-flow  engine). 

14.  The  critical  back  pressure  is  determined  by  the  initial  pressure,  the  clea- 
rance volume,  the  mean  effective  pressure,   and  the  length  of  compression.    It 
increases  in  proportion  to  the  initial  pressure  and  faster  tlmn  proportionally  to 


43 

the  clearance  volume;  it  first  increases  with  increasing  length  of  compression, 
then  decreases  with  increasing  length  of  compression  and  increasing  mean  effec- 
tive pressure.  It  is  zero  for  initial  pressure  =  zero,  clearance  volume  =  zero, 
length  of  compression  =  zero,  and  attains  a  maximum  for  mean  effective  pres- 
sure =  zero.  The  clearance  volume  has  by  far  the  greatest  influence,  the  pressure 
range  has  less,  and  the  length  of  compression  and  mean  effective  pressure  have 
the  least. 

15.  In  compound  engines  the  low  pressure  cylinder  determines  the  critical 
back  pressure.    Under  the  same  conditions  compounding  reduces  the  critical  back 
pressure  corresponding  to  the  lower  initial  pressure  of  the  low  pressure  cylinder. 

16.  It  is  not  so  important  merely  to  achieve  low  critical  back  pressure  alone 
as  it  is  to  obtain  simultaneously  small  volume  losses  and  low  critical  back  pres- 
sure.   The  volume  loss  may  be  very  large  and  the  critical  back  pressure  may  still 
be  zero.    Of  all  single  cylinder  condensing  engines,   the  single-beat  poppet  valve 
una-flow  condensing  engine  has  at  the  same  time  the  smallest  volume  loss  and 
a  critical  back  pressure  which  is  far  below  anything  that  can  be  reached  even 
with  the  most  modern  condensing  equipment,  mainly  on  account  of  its  small  clea- 
rance of  less  than  1%  and  its  favorable  length  of  compression. 


44 


2b.   Additional  Clearance  Space. 

Practically  all  condensing  una-flow  engines  must  be  able  to  run  non-condensing. 
In  case  of  breakdown  of  the  condenser,  lack  of  cooling  water,  or  during  the  winter 
months  when  the  exhaust  steam  is  used  for  heating  purposes,  the  engine  must 
be  capable  of  operation  with  either  atmospheric  or  higher  back  pressures.  The 


Fig.  1. 


Fig.  2. 

simplest  way  of  accomplishing  this  purpose  is  the  provision  of  an  additional  clea- 
rance space.  (See  Figs.  1  and  2.)  The  amount  of  additional  clearance  depends 
upon  the  initial  and  back  pressure.  If  the  latter  is  for  instance  1  at.  abs.,  the 
initial  pressure  13  at.  abs.,  and  the  clearance  for  condensing  operation  1.5%, 
then  the  additional  clearance  should  be  14.75%,  according  to  the  tables  to  be 
given  later.  At  the  same  time  this  increased  clearance  will  cause  a  lengthening 


45 


of  the  cut-off  from  8  to  12%  for  the  same  output  with  non-condensing  operation 
(Fig.  3).  The  drop  in  pressure  at  the  end  of  expansion  amounts  to  0.8  at.  for  con- 
densing and  1.0  at.  for  non-condensing  operation.  It  is  found  that  for  other  initial 
pressures,  with  approximately  the  same  drops  of  pressure  (0.8  or  1  at.  abs.)  at  the 
end  of  expansion,  the  mean  effective  pressures  produced  are  about  equal.  In 
the  previous  chapter  it  was  demonstrated  that  the  mean  specific  volume  loss  for 
saturated  steam  of  13  at.  abs.  and  non-condensing  operation  was  0.0918  kg/HP- 
hour  and  per  1%  clearance,  and  0.072  kg/HP-hour,  and  per  1%  clearance  for 
superheated  steam  of  300°  G,  the  other 
data  being  the  same.  For  14.75%  clea- 
rance the  total  losses  amount  to  1 .35  and 
1.06  kg  respectively  in  the  two  cases. 
The  general  adoption  of  the  additional 
clearance  space  despite  this  considerable 
increase  in  steam  consumption  is  due 
firstly  to  its  simplicity,  and  secondly 
to  qualities  which  tend  partly  to  coun- 
teract this  heavy  loss.  The  effect  on 
the  overall  economy  is  negligible  if  an 
engine  operates  with  additional  clear- 
ance only  for  several  days  or  hours 
during  the  course  of  a  year.  Fig.  3 
also  indicates  that  although  the  drop 
in  pressure  at  the  end  of  expansion 

is  higher  when  using  the  additional  clearance  space,  the  loss  due  to  incomplete 
expansion  is  less;  and  the  condensing  cylinder  being  rather  large  for  non- 
condensing  operation  becomes  more  or  less  adapted  to  this  condition.  The 
additional  clearance  also  preserves  the  una-flow  principle,  including  the  series 
arrangement  of  live  steam  space,  inlet  valve,  piston  and  exhaust,  which  is 
such  a  valuable  feature  of  the  una-flow  engine.  Although  it  is  possible  to  use 
auxiliary  exhaust  valves  instead  of  additional  clearance,  and  these  valves  being 
relieved  of  pressure  at  the  time  of  opening  can  be  of  single  beat  or  annular  con- 
struction, no  joint  at  all  being  preferable  to  even  a  tight  joint  or  seat. 

Care  must  be  taken  that  the  clearance  valves  which  control  the  additional 
clearance  pocket  do  not  materially  add  to  the  cylinder  clearance  for  condensing 
operation  (Figs.  4  and  5).  In  this  respect  it  is  advantageous  to  provide  the  clea- 
rance valves  with  projections  which  fill  up  the  space  between  valve  seat  and  cylin- 
der surface.  The  valve  area  of  the  clearance  valves  must  be  large  enough  to  avoid 
throttling  during  expansion  and  compression.  It  is  also  advisable  to  arrange 
the  additional  clearance  so  that  it  will  act  as  a  kind  of  heat  insulator  when  the 
engine  is  running  condensing,  which  is  especially  of  .value  for  the  crank  end  of  the 
cylinder. 

The  clearance  valve  may  also  be  designed  in  the  form  of  a  spring  loaded  safety 
valve,  but  then  the  above  mentioned  projections  cannot  be  used.  The  safety 
valve  action  of  the  clearance  valve  is  unnecessary  when  the  main  steam  valve 
is  not,  or  only  partly  balanced  so  that  it  can  act  as  a  safety  valve.  The  "safety" 


inlet  valve  is  preferable  to  the  "safety"  clearance  valve  since  its  weight  and  spring 
load  are  less.  The  inlet  valve  is  designed  for  high  speed  and  held  closed  by  steam 
pressure,  its  spring  being  only  strong  enough  to  overcome  inertia.  The  clearance 
valve  on  the  other  hand  is  heavy  and  its  spring  has  to  overcome  the  total  steam 
pressure.  In  case  of  sudden  failure  of  the  vacuum  this  heavy  spring  load  combined 
with  the  great  weight  of  the  valve  cause  an  objectionable  hammering,  which  can 
only  be  stopped  by  screwing  the  valves  back. 

In  Fig.  6  is  reproduced  a  diagram  such  as  is  obtained  from  a  una-flow  engine 
running  non-condensing  and  fitted  with  auxiliary  exhaust  valves,  the  clearance 
being  the  same  (1%%)  as  for  condensing  operation.  It  is  evident  that  the  ratio 


Fig.  4. 


Fig.  5. 


of  expansion  is  too  high.  A  construction  of  this  kind  is  shown  in  Fig.  7,  having 
the  auxiliary  exhaust  valves  arranged  in  the  cylinder  heads.  The  increase  in  clea- 
rance volume  due  to  these  valves  was  not  taken  into  consideration  in  the  diagram 
of  Fig.  6.  The  diagram  indicates  that,  assuming  the  same  mean  effective  pressure, 
the  expansion  line  reaches  the  back  pressure  while  the  piston  uncovers  the  exhaust 
ports.  The  loss  due  to  incomplete  expansion  is  zero.  The  shape  of  the  diagram 
indicates  the  counter-flow  action  and  proves  that  the  cylinder  is  too  large  for 
non-condensing  operation,  especially  for  smaller  loads,  when  the  toe  will  change 
into  a  loop.  This  produces  a  backflow  of  exhaust  steam  into  the  cylinder  and  a 
corresponding  increase  in  condensation  losses.  For  loads  higher  than  normal  the 
exhaust  action  will  be  partly  una-flow  and  partly  counterflow.  The  loop  at  the 
end  of  expansion  cannot  occur  in  engines  fitted  with  additional  clearance  spaces. 


47 


The  worst  feature  of  auxiliary  exhaust  valves,  however,  is  their  detrimental 
effect  upon  the  condensing  operation  of  the  engine.  They  increase  the  clearance 
space  and  the  harmful  surfaces  as  well 
as  the  possibility  of  leakage,  and  sacrifice 
the  very  valuable  series  arrangement  of 
live  steam  space,  inlet  valve,  piston  and 
exhaust.  For  1%  increase  in  clearance 
volume,  an  additional  steam  consumption 
of  0.12  kg/IHP-hour  may  be  expected, 
superheated  steam  being  assumed.  This 
figure  does  not  include  the  effect  of  leak- 
age and  the  surface  losses  caused  by  the 
valves  and  their  pockets,  nor  additional 
losses  due  to  the  operation  of  these  valves 
while  the  engine  is  running  condensing. 
It  is  advisable  to  keep  these  valves  in 

operation  even  while  running  condensing,  Fi     6 

since  they   are  liable   to   stick   after  re- 
maining out  of  use  for  some  time.    If  the  auxiliary  exhaust  valves  shorten  the  length 
of  compression  also  for  condensing  operation,  a  larger  volume  loss  results,  because 


48 


an  increase  in  clearance  necessitates  a  corresponding  lengthening  of  the  compres- 
sion. The  bad  effect  upon  condensing  operation  appears  all  the  more  objectionable 
since  it  occurs  during  the  whole  working  period;  while  on  the  other  hand,  for  short 
periods  of  non-condensing  service  even  a  considerable  increase  in  steam  consump- 
tion due  to  additional  clearance  could  be  tolerated.  For  long  periods  of  non-con- 
densing operation,  as  for  instance  during  the  winter,  single-beat  auxiliary  exhaust 
valves  are  preferable  to  additional  clearance. 

When  auxiliary  exhaust  valves  are  used,  they  are  usually  placed  below  the 
engine  room  floor  level,  which  renders  their  attendance  difficult  and  the  arrange- 
ment of  the  piping  awkward. 

The  amount  of  the  necessary  clearance  is  determined  by  the  following  rule: 
for  a  given  length  of  compression,  mean  effective  pressure,  initial  pressure  and 
back  pressure,  in  order  to  keep  the  volume  loss  as  small  as  possible,  the  clearance 
volume  should  be  made  large  enough  to  produce  equal  variation  of  pressure  during 
expansion  and  compression  (see  chapter  on  volume  loss).  On  an  average,  the  ter- 
minal expansion  pressure  for  non-condensing  operation  may  be  taken  as  1  at. 
gage,  and  this  implies  a  terminal  compression  pressure  of  1  at.  below  initial 
pressure. 

The  following  table  gives  the  total  amount  of  clearance  volume  required, 
when  operating  non-condensing,  for  90%  length  of  compression,  starting  with 
a  pressure  of  1.03  at.  abs.  and  ending  1  at.  below  initial  pressure,  with  adiabatic 
compression  and  saturated  dry  steam.  The  figures  are  based  on  the  latest 
Mollier  chart. 


Initial  Pressure 

8 

9 

10 

11 

12 

13 

14 

15 

16 

at.  abs. 

Total  Clearance 

27.9 

23.8 

21.3 

19.2 

17.6 

1625 

15.15 

142 

13.4 

O/ 

/o 

49 


3  a.  Losses  due  to  Throttling. 

Throttling  is  a  change  of  state  in  which  the  total  heat  remains  constant,  the 
effect  of  which  is  to  diminish  the  amount  of  heat  and  the  pressure  difference  avail- 
able for  utilization  between  boiler  and  condenser.  Losses  due  to  throttling  may 
occur  in  the  superheater,  steam  main  from  superheater  to  the  engine,  stop  valves, 
inlet  valves,  piston-controlled  exhaust  ports  or  exhaust  valves,  and  in  the  exhaust 
pipe  between  engine  and  condenser  (Fig.  1). 

The  losses  due  to  throttling  occuring  in  the  superheater,  steam  pipe,  stop 
valves,  and  inlet  valves  may  be  partly  regained  in  connection  with  the  subsequent 
expansion,  although  the  greater  part  is  lost.  The  percentage  of  this  regain  depends 


W 


Fig.  1. 


Fig.  2. 


upon  the  extent  of  the  expansion.  The  temperature-entropy  diagram  in  Fig.  2 
shows  the  conversion  of  a  certain  quantity  of  heat  at  high  temperature  and  small 
entropy  into  an  equal  quantity  at  lower  temperature  and  larger  entropy.  The 
change  is  represented  by  the  area  GCDQKHG  and  is  equal  in  area  to  the  strip 
QKLWQ  which  results  from  increasing  the  entropy.  The  part  KNVJK,  falling 
within  the  area  of  expansion  will  be  regained,  while  the  part  NLWVN  below 
the  line  of  terminal  expansion  pressure  is  definitely  lost.  Throttling  losses  occuring 
in  the  exhaust  valves  or  exhaust  pipe  are  irretrievable,  for  which  reason  they  must 
be  restricted  to  the  smallest  possible  amount.  They  are  especially  harmful  because 

S/ump/,  The  una-flow  steam  engine.  4 


50 

their  effect  extends  through  the  whole  compression  stroke.    (See  chapter  on  the 
relation  of  the  una-flow  engine  and  the  condenser.) 

The  heat  losses  and  throttling  losses  in  the  steam  pipe  cannot  be  separated 
and  are  usually  combined  in  one  figure;  a  loss  of  0.5  to  1.0%  is  considered  an 
average  and  corresponds  to  a  steam  velocity  of  about  40  to  50  m/sec,  calculated 

on   the   total  amount   of  steam   flowing 
through. 

The  throttling  losses  occuring  between 
the  stages  of  multistage  engines  are  eli- 
minated in  the  una-flow  engine. 

In  order  to  estimate  the  throttling 
losses  in  the  inlet  and  exhaust  valves  it 
is  necessary  to  know  the  relation  between 
effect  (losses)  and  cause  (valve  area),  to 
FA which  the  following  calculations  refer. 


f 


^5 


A 


P: 


Fig.  3. 


Determination  of  inlet  valve  areas. 

In  Fig.  3,  pt  and  vt  represent  the 
pressure  and  volume  of  the  steam  at  the 
dead  center  position  of  the  piston;  piston 
travel  x1  =  0  and  corresponding  crank 
angle  6^  =  0.  p2,  v2  are  the  pressure  and 
volume  at  the  point  of  valve  closing, 
for  a  piston  travel  =  x2  and  crank  angle 
=  (52.  p  and  v  are  the  pressure  and 
volume  at  any  intermediate  point  where 
the  piston  travel  =  x  and  crank  angle 

—  <5;  w  represents  the  velocity  of  the  entering  steam  corresponding  to  the  pre- 

vailing pressure  difference. 

F  is  the  valve  area  in  sqm, 
<p  the  velocity  coefficient, 
y,  v  the  specific  weight  and  specific  volume, 

t  the  duration  of  steam  admission  corresponding  to  piston  travel  #2, 
Q,  V  the  weight  and  volume  of  steam  contained  in  the  cylinder  correspond- 
ing to  piston  travel  x 

dQ  =  (p-W'F-ydt    .........     (1) 


The  change  of  state  during  throttling  is  represented  approximately  by  the 
equation 

=  Pv  =  Pzvz  =  c 


- 
~  P 


51 


n  •  360  ' 


6  Al 


dQ= 


v.w.  F-  f  dt=  I-  d-P  + 

V         P 

_d^         .W.F.1  .dt 
60  •  (50  d  d 


0 

dV 


Z)  and  H  are  the  cylinder  diameter  and  stroke  respectively  in  meters  ;  s  =  clea- 
rance volume  in  %  of  stroke.    Engine  constant  =  D2  •  H  •  n  =  A 

~ 
dp  4 

P 


d  (x  +  s)         <p>  w-F'dt 


n 


D"-  •  H  (x  +  s)          :-Dz-H(x 


dx          0,2125  •  y  w-F'dd 
x-{-  s  A  (x  +  s) 


(2) 


For  a  small  drop  in  pressure  it  may  be  assumed  with  sufficient  accuracy  that 

w  =  i  *  g  (Pi  —  P)  •  »! 

In  order  to  facilitate  the  calculations,  the  admission  line  or  rather  the  curve 
representing  the  change  in  pressure  plotted  against  crank  angle  or  time  will  be 
replaced  by  a  parabola.  It  will  be  shown  later  that  this  assumption  is  admissible 
if  the  object  of  the  calculation  is  not  the  shape  of  the  admission  line  but  the  final 
drop  of  pressure  at  the  end  of  admission.  This  final  pressure  drop,  however,  is  to 
form  the  basis  of  the  determination  of  the  inlet  valve  areas.  Therefore  we  may 
write 

pl  —  p  =  a  d-  and  p1  —  pz  =  a  dz2 
or  2 


and  therefore 


combining  equations  (3)  and  (2) 

dp         dx          0.2125- 

+ 


Jdp    .    C    dx 
P    hj  s-h 


2  .9) 


p1  (p  —  pt)  •  F-6-dd 


0.2125  .  <pi[  2  g- DI jjoj  —  p2)  -CF.  d-dd 

A  -  <52  J     x  -j-  s 

o 


4* 


(3) 


52 


F-d.dd 
x  -h  s 


A  <32  •  2.444 


log 


2  s) 


10 


(4) 


Assuming  the  admission,  clearance  volume,  initial  pressure,  initial  tempera- 
ture, pressure  drop,  velocity  coefficient  and  the  engine  constant  to  be  known 

quantities,  then  the  right  hand  side  of  equation 
4  reduces  to  a  numerical  value.  If,  for  ^4=1, 
this  value  is  C.  then  for  any  other  value  of  A 
the  result  is  CA,  or 


fF'd-dd 

J      *  +  s 


=  AC 


and 


/-» __ 


V  2,444 


~  10g 


io 


Fig.  4. 


If  the  driving  element  is  an  eccentric  or 
a  crank,  and  if  the  valve  seats  are  flat,  then 
we  may  write  h  =  a  -  Amax  and  correspondingly 
F  =  a  •  Fmax  (Fig.  4). 

s 

Multiplying  a  •  Fmax  by  -      -  and  plotting 

X  ~Y~  S 


this  product  on  a  diagram  against  <5  as  abscissae, 
a  curve  is  obtained  as  shown  in  Fig.  4,  the  area  of  which  is 


=      a-F* 


Accordingly 


d-dd 


-  \\T\1 0  T* O          /-?    •         I 

~B~*  J  ~^+ 


d  6 


•If 


I}    s   and 


as  well  as  the  pressure  drop  p{  —  p2  are  known,  then  .Fmax  can  be  easily  calculated. 

Permissible  average  values. 

For  most  una-flow  condensing  engines  the  following  values  can  be  assumed: 
Initial  pressure  px          =  13  at.  abs. 
Steam  temperature  t±  =  300°  Centigrade. 
Specific  volume  v^         =  0.2  cbm/kg. 
Clearance  Volume  s      =3%  (double  beat-valves). 

Velocity  Coefficient  9?  =  0.6  for  double-beat  valves,  piston  valves  and  slide  valves. 
Velocity  coefficient  9?    =  0.8  to  0.9  for  single-beat  and  Corliss  valves  (the  higher 

figure  for  machined  ports  and  passages). 
Lead  of  the  steam  valve  =  0. 
With  the  above  values  and  9?  =  0.6  (double-beat  valve), 


/ 130  000  —  p2 


"7800" 


53 


Values   of  C,  calculated   by  means  of   this   formula   for  various  admissions 
#2   and  various  pressure    drops    at   the  end    of  admission  are  plotted  in  Fig.  5. 
For  any  cut-off  and  pressure  drop,  the  corresponding  value  of  C  may  be  read  off. 
Multiplying  this  by  the  engine  constant  A,   the 
right  hand  side    of  equation  4    is   disposed   of. 


side  I 

o 


T-T          5          j      5 

—  '-  ----- 

x  +  s 


may 


be   inte- 


grated  graphically  for  each  separate  case. 

The  first  step  is  to  lay  out  a  curve  showing 
the  valve  lifts  h  plotted  against  piston  travel 
for  each  cut-off.  The  corresponding  valve  areas 
are  F  =  b  -  h  (b  =  width  of  port)  for  slide 
valves,  F  =  n  -  d  •  h  for  single-beat  valves,  and 
F  =  1  •  7i  •  d  •  h  for  double-beat  valves,  (d  = 
smallest  valve  seat  diameter).  For  each  area  F 
the  corresponding  values  of  d  and  x  and  there- 

F  •  ft 
fore  of  the  expression       .    |  -  are  easily  obtained. 

'JU  I  O 

For  example,   the  piston  travel  is  30%   for  a 
crank   angle   of  66,4°   and   since  the  clearance 

d  66.4° 

volume  s   is  to  be   3%   -^+7  =  =  0.3 +  0.03' 

s 

Plotting         -  as  ordinates  against  d  as  abscissae, 


X  -j-  S 

then  a  narrow  strip  of  the  enclosed  area  on  the 

F'd-dd 
base    do   represents   the    expression    — — — 

If   the  curve  commences  at  d  =  0,    then   the 
enclosed  area  up  to  any  value  <52,  measured  by 

means  of  a  planimeter,  represents  the  expression 
'J. 
fiF-d-dd 


Z^io 


I 


10 


30 


t 


w 


Fig.  5. 


The  effect  of  the  fraction 


x  + 


is  of  interest  (Fig.  7).    It  increases  rapidly 


at  first  until  it  attains  a  maximum  for  a  crank  angle  of  d  =  20°,  after  which 
it  gradually  decreases.  This  curve  is  obtained  by  means  of  the  curve  Fig.  6,  which 
shows  the  relation  between  d  and  x. 

It  will  be  remembered  that  a  parabola  was  assumed  to  represent  the  admis- 
sion line  on  a  time  basis  and  this  parabola  of  course  presumes  a  certain  valve 
lift  or  valve  area  curve.  This  F  curve  could  be  developed  point  by  point  from 
the  assumed  parabola  and  naturally  would  differ  from  an  F  curve  based  on 
an  eccentric  circle.  The  latter  curve,  however,  may  be  used  in  this  case  for 


the   solution   of  the  expression 


, 

CF- 
\ 

J    z 


since,   as   will  be   shown  later,   both 


54 

kinds   of   curves   result  in    approximately   the    same  pressure   drop   at   the   end 
of  admission. 

If  a  curve  representing  valve  lifts,  as  produced  by  an  eccentric  gear,  is 
plotted  against  the  crank  angle,  then  each  valve  lift  can  be  expressed  in  % 
of  the  maximum  lift  (Fig.  8).  This 
percentage  is  given  at  5  points  and 
remains  constant  for  every  cut-off  and 

100* 


It 


XL 


60. 


_ikL 


M_ 


tt 


&<L 


J3S- 


Fig.  6. 


Fig.  7. 


the    same    divisions.     Therefore    the    valve    lift    curve    in    Fig.  9  may   be    con 
sidered  a  standard   for   every  cut-off;   or   in  other  words,  considering  the  maxi 


Fig.  8. 


55 


mum  valve  area  equal  to  unity,  then  the  figures  written  at  the  division 
lines  indicate  the  corresponding  valve  openings.  It  is  assumed  that  the  valve 
has  no  lead  and  does  not  overrun  the  port.  For  example,  after  1/8  and  7/s  of  the 
admission  time  have  elapsed,  the  valve  opening  for  any  cut-off  is  F  =  0.43  jPmax, 
after  %  and  %  time  F  =  0.75  ^max  and  after  3/s  and  5/8  time  F  =  0.94  .Fmax-  Each 

C 

division  also  corresponds  to  a  certain  value  of  — . — , 

X -{-  S 

which  of  course  varies  with  the  cut-off.  This  value 
may  be  combined  with  the  above  fractions  into  a 
constant  B  given  by  the  equation 


dd 


=  C 


f*K 


3,0 


150QO. 


i&m 


O          hO 


o£ 


ttftj 


Fig.  9. 


Fig.  10. 


or 


lx=  -~'.    This  equation  is  plotted  in  Fig.  10.    For  any  cut-off  the  value  of 

Fmax  is  obtained  by  taking  the  corresponding  value  of  B  from  Fig.  10  and  dividing 
it  into  the  value  of  C,  derived  from  Fig.  5  for  the  same  cut-off  and  some  parti- 
cular pressure  drop,  ^4.  =  1,  99  —  0.6,  s  =  0.03,  p1  =  13  at.  abs.,  ^  =  300°  C, 
direct  eccentric  drive  and  no  cam  mechanism  being  assumed.  The  values  of  Fmax 
obtained  in  this  manner  are  given  in  the  following  table  and  plotted  in  Fig.  11. 


h,= 

0,5  at 

1  at 

2  at 

3  at 

4  at 

5  at 

6  at 

x2=     5% 

1.84  cm2 

1.21  cm8 

0.72  cm2 

0.44  cm2 

0.27  cm2 

0.12  cm2 

0.02  cm2 

*2  =  10% 

2.67     „ 

1.8      „ 

1.13    „ 

0.82    „ 

0.61     „ 

0.44    „ 

0.28    „ 

*2  =  20% 

4.08    „ 

2.82     „ 

1.85    r, 

1.41     „ 

1.13    „ 

0.90    „ 

O.'i. 

x2  =  30% 

5.36     „ 

3.67     „ 

2.47    „ 

1.90    „ 

1.57     „ 

1.27    „ 

1.06    „ 

xz  =  40% 

6.34    „ 

4.48    „ 

3.02    „ 

2.35    „ 

1.91     „ 

1.59    „ 

1.33     „ 

xz  =  50°/0 

7.50    „ 

5.20    „ 

3.52     „ 

2.76    „ 

2.24    „ 

1.89    „ 

1.59     „ 

For  any  other  engine  constant  A^=D2H  •  n,  this  value  ofFmax  must  be  multi- 

C 
plied  by  A,  D  and  H  being  measured  in  meters.   Then  Fmax  =  A  -  -.  If  the  valve 


56 


is  operated  by  means  of  a  cam  then    the  expression 


<>t 

(   F  •  6- 

*  + 


dd 


57 


has  to  be 


diminished  by  a  certain  percentage  according  to  the  cam  profile.    It  must  also  be 


10" 


i\ 


\ 


ft  8 


Fig.  12. 


considered  whether  the  valve  remains  stationary  during  part   of  the  time  it  is 
open.    (See  Figs.  12  and  13.) 

In  case  the  steam  valve  has  a  certain  amount  of  lead  (Fig.  14),  the  integra- 
tion of  the  .F  curve  must  apply  only  to  the  area  after  the  dead  center,  since  the 


Fig.  14. 


Fig.  15. 


purpose  of  lead  is  merely  to  fill  up  the  clearance  space,  a  condition  which  was 

assumed  from  the  start. 

CL  =  ^2 
C»       Vi' 


The  values  of  C  are  inversely  proportional  to  (p,  and 


Further,  ^maXj  = 


-         .  nr 

—  ~         ^        or 


9>1 


Having  found  .Fmax  for  a  certain  cut-off  and  pressure  drop,  the  question  arises 
of  the  pressure  drop  for  different  cut-offs.  For  instance,  at  12.5%  cut-off  and 
hz=  2  at.  pressure  drop,  the  value  of  .Fmax  as  taken  from  the  curves  in  Fig.  11  is 
1.325  sqcm.  Assuming  a  direct  drive  from  the  eccentric,  the  values  obtained  for 


58 

the  pressure  drop  for  different  cut-offs  are  given  in  the  following  table  and  plotted 

in  Fig.  15. 

*2  =  5°/0     10%     12.50/0     15%    20%    30%    40%    50% 
fc2  =  2.55     2.2         2.0        1.8       1.55      1.15      0.8       0.7. 

Fig.  15  also  contains  two  more  curves  for  higher  pressure  drops  during  normal 
admission.  These  curves  show  that,  if  for  a  condensing  una-flow  engine  the  pres- 
sure drop  is  normal  for  rated  cut-off,  then  a  larger  cut-off  will  show  a  smaller  pres- 
sure drop.  For  a  small  cut-off  the  pressure  drop  increases  still  further,  while  in 
case  of  a  higher  pressure  drop  at  normal  cut-off  this  gradually  tends  to  be  a  maxi- 
mum. For  condensing  una-flow  engines  it  is  sufficient  to  calculate  ^max  f°r  normal 
cut-off. 

Example. 

The  necessary  valve  area  of  a  single-beat  valve  for  a  una-flow  stationary 
engine  is  to  be  calculated  for  the  following  conditions.  Cylinder  diameter  D  =  0.4  m, 
stroke  H  =  0.5  m,  r  •  p  •  m  =  150,  steam  pressure  p±  =  13  at.  abs.,  steam  tem- 
perature ^  =  300°  centigrade,  assumed  pressure  drop  ='2  at.  for  12.5%  valve 
gear  cut-off,  or  z2  =  0.125  and  <52  =  41.4°,  <p  =  0.8,  5  =  0.01. 

41.4-2.44  110000  (0.125  -f  2  .  0.01) 

°gl° 


1/0.2  (130000  —  110000) .  0.8  130000  •  2  •  0.01 

Engine  constant  A  =  0.42  •  0.5  •  150  =  12. 

The  large  area  in  Fig.  13  represents  the  value  of  I  -    — ~      for  direct  eccentric 

J      x-{-s 

0 

drive,  while  the  smaller  shaded  area  =  2230  sqmm  gives  the  same  integral  when 
a  cam  or  rolling  lever  is  used.  If  the  scale  of  the  abscissae  is  1  mm  =  0.5°  and 
that  of  the  ordinates  is  1  mm  =  10  Fm!LK,  then 

2230  •  0.5  •  10  Fmax  =  11150  ^max  =  B  -  FmZK 

anc*  r        1010 

1.26  ==0>001356sqm2:=:13i56sqcm2> 


If  the  maximum  valve  lift  for  the  assumed  valve  gear  cut-off  of  12.5%  is 
equal  to  Vio  tne  valve  diameter,  then.Fmax  =  n  -  d  •  0.1  •  d  =  13.56  sqcm,  d  =  6.6  cm, 
and  Amax  =  0.66cm.  The  common  empirical  formula,  based  on  mean  piston  velo- 

0  •  cm 
city  would  give  a  steam  velocity  wm  =  -•-  —  m  =  92  m/sec.  These  calculated  values 


of  F'max  and  ftmax  can  be  realized  without  difficulty  if  the  single-beat  valve  is  ope- 
rated by  a  lay  shaft  gear  running  at  twice  the  engine  speed  (See  final  chapter). 

Permissibility  of  the  use  of  the  parabola. 

It  is  still  to  be  proved  that  the  actual  diagram  admission  line  plotted  against 
crank  angle  or  time  may  rightly  be  replaced  by  a  parabola.  For  this  purpose  a 
diagram  is  laid  out  with  the  crank  angles  d  as  abscissae  and  the  valve  openings 
as  ordinates.  This  results  in  a  curve  such  as  that  shown  in  Fig.  16.  The  total 
crank  angle  is  now  divided  into  a  large  number  of  parts  or  intervals  and  for  each 


59 


part  the  mean  valve  area  is  determined  (F^  F2,  Fs  etc.).    It  is  assumed  that  p  •  v 
=  const.  In  the  above  example  it  was  assumed  that  p1  =  130000  kg/sqin,  ^  =  300°, 

i>!  =  0.2  cbm/kg  and  /v  ^  =  26000. 

y 
Volume  of  clearance  space  Fj  =  s\  weight  of  steam  Ql  =  -— ;  stroke  volume 

0  -f  w2         w2  ^   * 

~~2~       v    2 


=     H\   «'«.= 


=  O.S-  dt  = 


for    i    intervals.      The 


n  •  360  •  i 
piston  is  first  considered  to  be  moved  forward  a  distance  corresponding  to  the 

V1               26000 
first  interval  without   admission  of  steam,   so  that  p2  =  pr     y  ',    v2  =     ; 

;>2  —         and  ym  =  -^——2> 

('o  — 

These  values  represent  the  state  of  the  steam  at  the  end  of  the  first  interval, 
produced  by  expansion  only,  without  the  admission  of  live  steam.  wm  and  ym 
being  also  known,  it  is  now  possible  to  calculate  the  additional  weight  of  steam 


Fig.  16. 

admitted,  which  is  dQ  =  <p  •  wm  •  Fl  •  ym  •  dt.  Then  at  the  end  of  the  first  interval 
the  cylinder  contains  the  amount  Q2  =  Qt  -f  dQ  with  a  corresponding  volume  of 
V*  =  VH(s  +  dxj  in  which  dx±  represents  the  distance  the  piston  has  traveled 
during  the  time  dt.  This  gives  the  actual  specific  volume  of  the  steam  at  the  end 

V2                                    ,      26000 
oi  the  first  interval  v  2  =  n    and  the  pressure  p2  = -f — • 

\L  2  ^2 

For  the  second  interval,  the  piston  is  again  considered  moved  forward  a  dis- 
tance corresponding  to  the  second  interval,  without  steam  admission.  Starting 
with  the  state  of  the  steam  at  the  end  of  the  first  interval,  v3  and  /?3  are  calcu- 
lated, and  the  same  procedure  is  repeated  as  before.  In  this  manner  the  curves 
shown  in  Fig.  17  were  obtained,  wherein  the  parabolas  are  the  dotted  lines.  They 
are  plotted  for  25%  and  40%  valve  gear  cut-off,  both  for  the  true  admission  line 
and  the  substitute  parabola.  The  result  is  an  approximately  equal  pressure  drop 
at  the  end  of  admission  in  both  cases.  If  shorter  intervals  had  been  used  the 
results  probably  would  have  agreed  still  more  closely.  Although  the  admission 
line  based  on  the  parabola  differs  considerably  from  the  shape  of  the  actual  ad- 
mission line,  it  gives  practically  the  same  final  pressure  drop;  and  if  the  valve  area 
is  to  be  based  on  the  latter  the  parabola  may  be  used  to  advantage.  It  is  assumed 
of  course  that  the  pressure  at  the  admission  valve  is  constant,  which  can  be  insured 


60 

by  the  proper  amount  of  steam  storage  space  around  the  valve  and  by  steam 
pipes  of  sufficient  size.  If  this  is  not  provided  then  the  total  pressure  drop  at  the 
end  of  admission  increases  by  the  amount  of  pressure  loss  at  the  inlet  valve. 

The  drop  of  pressure  at  the  end  of  admission  corresponds  to  a  certain  thrott- 
ling loss.    The  initial  state  of  the  steam  p^  yx  changes  into  p%,  u2  at  the  end  of 


u 


-to 


£ 
5  A 


/) 


/.als. 


Z6 


39 


( *»  ?  /<  / 

5fe°  g?°^5-' 


k 


^  5 


78 


f  cWil/. 


/3    <ZO 


Ctfe. 


i°/ 

/o 


*f 


V£)% 


Fig.  17. 

admission,  and  the  subsequent  expansion  allows  a  part  of  this  loss  to  be  regained. 
The  amount  of  this  regain  as  well  as  the  total  final  loss  can  be  found  from  the 
temperature-entropy  diagram  (Fig.  2). 


Calculation  of  Exhaust  Port  Areas. 

Throttling  losses  in  the  exhaust  ports  or  valve  and  in  the  exhaust  pipe  are 
irretrievable  because  there  is  no  subsequent  expansion  by  which  they  may  be 
regained.  The  dimensioning  of  the  exhaust  ports  therefore  requires  particular 
caution.  The  loss  in  the  diagram  due  to  incomplete  expansion  will  be  dealt  with 
later  on.  The  point  of  interest  at  this  time  is  the  exhaust  throttling  loss  caused 
by  insufficient  area  of  the  exhaust  passages,  which  makes  itself  noticeable  by  the 
rounding  off  of  the  end  of  the  diagram  as  well  as  a  narrow  strip  along  the  exhaust 
and  compression  lines.  The  latter  part  of  this  loss  is  especially  harmful;  it  may 
be  avoided  by  placing  the  condenser  close  to  the  cylinder,  and  making  the  con- 
nection between  them,  as  well  as  the  exhaust  ports,  of  ample  area  (See  pages 
68 — 69.).  Throttling  losses  in  the  exhaust  are  aminimum  if  the  exhaust  lead  is  made 
as  small  as  possible  and  the  pressures  are  completely  equalized  at  the  point  of 
exhaust  closure.  Perfect  equalization  of  pressure  is  most  important.  In  order 
to  reduce  the  exhaust  lead  to  a  minimum,  the  ports  must  occupy  as  much  as  pos- 
sible of  the  cylinder  circumference,  only  enough  of  the  material  being  left  to  take 
the  strain  due  to  the  piston  load.  This  applies  more  particularly  to  condensing 
engines,  while  for  atmospheric  exhaust  a  small  part  of  the  circumference  is  suffi- 


61 

cient.  Larger  exhaust  lead  requires  a  smaller  exhaust  port  area.  When  using 
large  exhaust  lead  and  atmospheric  exhaust,  one  exhaust  port  is  sufficient  under 
certain  conditions.  (See  chapter  on  Una-Flow  Locomotives.)  The  requirement 
of  perfect  pressure  equalization  can  be  satisfied  on  the  basis  of  the  following 

calculation: 

dQ  =  cp  •  w  •  F  •  y  '  dt, 

in  which  q>  =  coefficient  of  velocity,  w  =  the  velocity  of  the  exhaust  steam  in 
m/sec,  F  =  instantaneous  value  of  the  exhaust  port  area  in  sqm,  y  =  specific 
weight  and  t  =  duration  of  exhaust. 

Even  with  highly  superheated  live  steam,  the  exhaust  steam  of  condensing 
engines  is  always,  and  that  of  non-condensing  engines  in  most  cases,  saturated. 
The  change  of  state  within  the  cylinder  can  therefore  be  assumed  to  follow 
Mariotte's  law: 

/?!  •  t>j  =  pv  =  p2vz  =  const. 

Pii  yi5  7i  represents  the  state  of  the  steam  at  beginning  of  exhaust, 
p,  y,  y  the  state  of  the  steam  at  any  intermediate  point, 
Pzi  vzi  72  the   state  of  the  steam  in  the  exhaust  pipe   (condenser,    atmo- 
sphere,  etc.), 

Pe,   ve,  7e   the  state    of    steam   at    the    narrowest    place    of  exhaust  port. 
x  represents  the  piston  travel  in  %  of  the  stroke  measured  from  the 
admission  end. 

As  long  as  — -  <  0.577  is  pe  =  0.577  p.       The  value  0.577    remains  about 
P 

the  same  for  any  steam  wetness.  If  p  <  1.735  p2  then  pe  =  p2.  The  change 
of  the  cylinder  volume  during  exhaust  is  neglected.  Since  the  weight  of 
steam  present  in  the  cylinder  is  proportional  to  the  absolute  pressure, 

p  —  dp       Q  —  dQ      _       T.         ,n  „          , 

J-  =  -— PC — —  ;    Q  =V -y;  dQ  =  w-(o- F  •  ye-dt 

p  Q 

* 

dp d Q  y  •  (Q-F  •  ye •  d t  .. 

~p"~=~Q~         ~^Tr~ 

t  =  — — — -  ;     dt  = pr  (<5  =  crank  angle  corresponding 

*-360  »'6  to  time  t) 

dp        <p.a>-F-dd.7e       v_^.D2,Ht(x  ,  s}.    A  —  jp.n.j 
~Y  S-n-V-y  -4     D  'H'(x  +  s>>  A 


C*1      1)  •     •  .    —     i      _      I  w  \*s         j.  \fit      v  f  g  /O  \ 

IP"-      A.(X+S).Y~ 

Range  of  High  Pressures. 

In  this  case  /?>  pcr=  1.735  p2,  w  =  3.23  y  p  v  =  3.23  •  j/c .  For  practical 
purposes  it  is  sufficiently  exact  to  replace  the  variable  quantity  (x  +  .9)  of 
equation  2  by  the  constant  quantity  (5  +  1  —  0.5  a),  where  a  =  exhaust  lead  (the 
critical  pressure  is  reached  approximately  at  the  dead  center). 


62 


0.2124.  y./c-  3.23 
.(s  +  l  —  0.5  a).  1.62 


1.62 


,„ 


per  Ocr 

(Xp  _       0.423.  y.]/7     f 
J    p      ^(s  +  l_0.5)J 


Fig.  18. 

The    quantities    F    for    the    range    of    high 
pressures  may  be  plotted   against  <5   according  to   Fig.  18  and 


*r 

f  J 

0 

Then  loge 


px\_         0.423  .  y  •  f  c 
r  ^1  (s  +  1  —  0.5  a) 
A 


9? 


-  0.5  a)  .  5.45  -       loglo 

c  1.735/?2 


(3) 


in  m2,  ^4  in  m3/min,  c  in  kg/m2  •  m3/kg. 


For  condensing  una-flow  engines  with  double-beat  valves  the  average  clearance 
may  be  assumed  to  be  s  =  0,03  and  y  =  0.9  for  drilled  exhaust  ports  with  well 
rounded  edges;  also  A  =  1  in  m3/min  and  c  =  ip1v1  =  15000 


FmH  -  dcr  =  0.0495  (1.03  —  0.5  a)  Iog10 


in  m2     .     .     .     (4) 


For  non-condensing  una-flow  engines  also,  the  steam  is  in  most  cases  saturated 
at  exhaust  and  therefore 


FmH  dc>.  =  0.0457  (1.11  —  0.5  a)  loglo 

with  /?!»!  =  €  =  17500,  5  —  0.11,  9>  =  0.9,  ^4=1. 

' 


n 


.     .     .     (5) 


Range  of  Low  Pressures. 

If  p  has  been  reduced  to  pcr  =  1.735  /?2»  tnen  Pe  —  Pz  and 


w  = 


Therefore  — ~  — 


0.2124- 


I-5I  p  Jl  - 
"—1 


P  I 


•F-dd 


p  A  (x  +  s) 

Inserting  1+5  —  0.5  a  for  x  +  s  and  c  for    p.y    the    last    equation    may 
be  written  as  follows: 


63 


dp 


0.2124 


P' 


P. 


A  (s  +  1  —  0.5  a) 
0.942-  We 


j_       A  (5  +  1  —  0.5  a)  r  x- 


F. 


For  J/:  •  d  d  =  FmN  .  ((52  —  <5e(.)  will  be 

der 

FmN  (<52—  <U  =  •••  J-  1.063  (5  -f  1  —  0.5  a)  ~  ^  dp 


TJie  integral  of  this  equation  is  independent 
of  /?2?  wherefore   it   may  be  written 


dp 


1.735 


|  


1  \     * 


When  integrating  graphically  the  ordinate  for 
p  =  1  will  be  indefinite.  By  partial  inte- 
gration the  integral  may  be  written 

i  i 


Fig.  19. 


.  .  .(5a) 


1.735  1.735 


r  /  1  \    * 

7.38  —  16.8     yi—   —  *-irfp 
J  \.P/. 

1.735 

and  by  graphical  integration 

7.38  —  16.8-0.1315  =  5.17 

Consequently  Fmjv  (<32  —  ^)  =  —4.063  (s  +  1  —  0.5  a) ^4Z= 


1.895(5+1—  0.5  a) 


(6) 


with  .F  in  m2,   ^4   in  m3/min.,   c  in  kg/m2-m3/kg. 

For  A  =1  and  9?  =  0,9  will  be  in  condensing  una-flow  engines  (c  =  15000; 

FmN  (82—^)  =  0.0172(1.03  —  0.5  a)  in  m2 (7) 

in  non-condensing  una-flow  engines   (c  =  17500;  5  =  0.11) 

FmN(^  —  dcr)  =0.0159(1.11—  0.5)  in  m2 (8) 


64 

Summary. 

According  to  Fig.  19  the  range  of  high  and  low  pressures  may  be  combined 

thuS  '  IT  c  T-I  e  j-i 

I'm  '  02  =  rmH  '  Ocr  +  Fm  N   (Oz  6cr). 

and  inserting  the   values  from  equations  3  and  6 

A     5.45  /       Pl      \       A  1.895  . 

m  z  =  <p  '~jT    +    ~         }    gl°  (iT^S^r-!  +  -  -y^(5  +  i  —  o.oa) 

FmS2=A.S+i-^a  [1.895  +  5.45 
V  }  c 

For  A  =  1   and  <p  =  0.9  will  be  in  condensing  una-flow  engines  (c  =  15000; 

S  =  °'3)       Fm  =  i^_a^  [0 0172  +  0  04Q5 .  togi.(T^hr.)l  .  .  .  .  (9) 

in   non-condensing  una-flow  engines  (c  =  17500; 


°*  "  [0. 


F~  =    '  0.0159  +  0.0457  loglo  _       ....     (10) 


1   ,         0.5  —  a 
Neglecting  the  definite  length  of  the  connecting  rod,  cos  -^-A  =    —  c\~xT~ 

1  —  2  a,  wherby  <52  can  easily  be  calculated.  Equations  (9)  and  (10)  should  be 
used  only,  if  Pi>  ;  1,735  Pftypi  <  1-735  p2  ti-e-—  =i.6  )}  the  value  (5a)  should 

\  ^2  / 

be  integrated  in  smaller  limits  (i.  e.  1.6  and  1)  (5.  upper  corner  in  Fig.  20).  In 
accordance  herewith  the  second  values  of  the  brackets  of  equation  (9)  and  (10) 
should  be  neglected,  if  they  become  negative. 

An  examination  of  engines  with  piston-controlled  exhaust  ports  (not  overrun 
by  the  piston)  shows  that       r-,  * 

-^max          "max          &  .    •. 

j-,          —r  —  =  ^y  approximately. 

•F  m  ""m  * 

The  following  table  as  well  as  Fig.  20  gives  the  values  of  Fm  and  -Fmax  (diffe- 
rent scales)  for  the  exhaust  through  the  high  and  low  pressure  ranges  for  various 

values  of  exhaust  lead  a  and  pressure  ratios  —  ,  A  being  =1,  live  steam  pressure 

Pz 

pl=  13  at.  abs.,  ^  =  300°,  <p  =  0,9  and  s=  3%  or  11%  respectively  for  conden- 
sing and  non-condensing  engines.  If  the  piston  overruns  the  exhaust  ports  Fm 
should  be  used,  otherwise  Fm3Lji  should  be  taken.  The  insert  in  Fig.  20  represents 
exhaust  in  the  low  pressure  range  alone.  Furthermore  it  is  immaterial  whether 
the  exhaust  ports  are  round  or  square.  The  shape  of  the  exhaust  ports  has  only 
little  bearing  upon  the  shape  of  the  exhaust  line  and  no  bearing  upon  the  terminal 
pressure.  This  is  shown  more  clearly  in  Fig.  21,  in  which  the  dashed  curves  refer 
to  round  and  the  full  curves  to  square  exhaust  ports.  The  conditions  assumed 
are  10%  exhaust  lead,  terminal  expansion  pressure  1.2  at.  abs.  and  back  pressure 
0.05  at.  abs.  These  curves  show  the  slight  deviation  of  the  exhaust  lines,  which 
were  calculated  point  by  point  by  the  interval  method,  as  well  as  the  fact  that 
in  both  cases  the  same  terminal  pressure  is  reached. 


65 


bo 

E 


Stumpf,  The  una-flow  steam  engine. 


66 


Exhaust  in  high  and  low  pressure  range  for  condensing  Engines. 


p±- 

Pa 

5 

10 

15 

20 

25 

30 

35 

40 

45 

50 

55 

60 

a  =  5% 

Fm    =7.76  cm2 
Fmax  =  12.2cm2 

10.7 
16.8 

12.35 
19.4 

13.6 
21.4 

14.5 

22.8 

15.25 
24 

15.9 
25 

16.45 

25.8 

16.95 

26.8 

17.4 
27.3 

17.75 
27.9 

18.1 

28.4 

a  =  7,5% 

Fm    =6.  33  cm2 
Fmax=9.95cm2 

8.57 
13.45 

9.99 
15.7 

10.9 
17.1 

11.6 
18.2 

12.25 
1925 

12.75 
20 

13.2 

20.7 

13.6 
21.4 

13.95 
21.9 

14.25 
22.4 

14.55 

22.8 

a  =  10% 

Fm    =53   cm2 
^max^S.SScm2 

7.3 
11.45 

8.45 
1325 

9.3 
14.6 

9.9 
15.5 

10.45 
16.4 

10.9 
17.1 

11.3 
17.7 

11.6 
18.2 

11.9 
18.7 

12.15 
19.1 

12.4 
19.5 

a  =15% 

Fm  =4.  19  cm2 
Fmax=6.58cm2 

5.78 
9.08 

6.67 
10.45 

7.33 
11.5 

7.82 
12.25 

8.25 
12.95 

858 
13.45 

8.9 
13.95 

9.15 
14.35 

9.4 
14.75 

9.6 
15.05 

9.8 
15.4 

a  =  20% 

^m  =3.49  cm2 
Fmax=5.48cm2 

4.8 
7.54 

5.55 
8.7 

6.11 
9.6 

6.52 
10.2 

6.87 
10.8 

7.16 
11.25 

7.41 
11.65 

7.63 
11.95 

7.82 
12.25 

8 
12.55 

8.16 
12.8 

a  =  25% 

Fm  =3.01  cm2 
Fmax=  4.72cm2 

4.14 
6.5 

4.79 
7.53 

5.27 
8.27 

5.62 
8.82 

5.92 
9.3 

6.18 
9.7 

639 
10 

658 
10.35 

6.75 
10.6 

6.9 
10.8 

7.04 
11.05 

Example. 

The  exhaust  port  area  will  now  be  calculated  for  the  same  engine  for  which 
the  inlet  valve  area  was  determined  in  the  previous  chapter. 

As  before,  D  =  0.4  m,  H  =  0.5  m,  r  •  p  •  m  —  150,  steam  pressure  =  13  at. 
abs.,  steam  temperature  =  300°  G. 

1.  Steam  pressure  at  beginning  of  exhaust,  Pi  =  1,6  at.  abs.,  back  pressure 

n2  =  0.03  at.  abs.;  hence   —  =  77770=  53.4. 

p%       u.uo 

Also  a  =  10%,  and  A  =  12  m3/min. 
From  Fig.  20  for  A  =  1,  ^max  =19  sqcm. 
Therefore  for  A  =  12,  Fmax  =  228  sqcm. 


67 


Round  ports  with  a  diameter  of  d  =  0.1  •  H  =  5  cm  have  an  area  of  /  = 


228 
19^35 


=  11.71.  e.!2holes. 


=  19.635  sqcm,  and  therefore  the  number  required  i  = 

2.  For  a  =  5%   and  ^=53.4,    ^max 

P2 

=  334  sqcm,  and  the  number  of  ports  required 
(2.5  cm  diameter)  i  =  68.  This  number 
cannot  be  realized. 

3.  For  a  =  25%  and  ^  =  53.4,  >max 

Pz 
=  129  sqcm.  With  a  port  diameter  of  12.5  cm, 

the  area  of  one  port  would  be  122.7  sqcm 
and  .  a  single  port  would  be  almost  suffi- 
cient. 

4.  For  the  same  engine  running  non-con- 
densing, the  conditions  may  be  pt  =  3  at.  abs. 

and  p2  =  1  at.  abs.,    hence  —  =  3,  a =20%. 

Pz 
Then  according  to  Fig.  20,  -Fmax  =  48.2  sqcm, 

and  one  port  of  10cm  diameter  having  an  area 
of  78.5  sqcm  is  therefore  already  too  large. 

The    formula    commonly    used,    based  »         Fig.  22. 

upon   mean  piston  speed,    would  give  the 

velocities  w  =.  13.8  m/sec  in  case  1,   w  =  9.4  in  case  2,   w  —  24.3  in  case  3,  and 
w  =  65  in  case  4.    This  proves  the  inadequacy  of  this  formula. 

In  Fig.  22  is  shown  a  diagram  in  which  the  admission  and  exhaust  lines  were 
calculated  point  by  point  after  the  areas  for  inlet  and  exhaust  had  been  found 
by  the  above  method,  a  drop  of  pressure  at  the  inlet  from  13  to  11  at.  abs.  and 
at  the  exhaust  from  1.2  to  0.05  at.  abs.,  13%  valve  gear  cut-off  and  10%  exhaust 
lead  being  assumed. 


68 


3b.   The  Relation  of  the  Una-Flow  Engine 
to  the  Condenser. 

A  high  vacuum  is  of  great  advantage  to  the  operation  of  una-flow  engines. 
Fig.  1  shows  compression  curves  for  different  back  pressures  and  the  same  ter- 
minal pressure,  the  clearance  volumes  being  correspondingly  changed.  These 
curves  indicate  how  appreciably  the  diagram  area  increases  with  better  vacuum. 
At  the  same  time  it  is  possible  to  keep  the  compression  up  to  the  desired  value 
by  properly  proportioning  the  clearance  volume.  For  a  high  vacuum  the  clearance 


Fig.  i. 

volume  used  may  be  very  small.    The  limit  is  usually  determined  by  the  design, 
2%  being  considered  an  average  figure. 

The  duration  of  the  exhaust  of  a  una-flow  engine  with  10%  exhaust  lead 
and  90%  length  of  compression  is  only  about  one  half  of  the  time  available  for 
the  exhaust  of  a  corresponding  counterflow  engine.  The  working  steam  of  the 
una-flow  cylinder  must  therefore  be  exhausted  into  the  condenser  in  one-half  the 
time.  It  is  a  fact  that  in  the  usual  design  of  counterflow  engines  there  exists  a 
considerable  pressure  difference  between  the  interior  of  the  cylinder  and  the  con- 
denser, which  is  used  to  overcome  the  resistances  in  the  usually  too  narrow  ex- 
haust passages.  The  shortening  of  the  duration  of  the  exhaust  in  the  una-flow 
engine  is  all  the  more  a  reason  for  diminishing  to  the  utmost  the  resistance  between 
condenser  and  engine  cylinder,  and  this  can  be  accomplished  by  short  passages 


69 


of  large  area.  Furthermore,  the  exhaust  port  area  of  the  una-flow  cylinder  can 
easily  be  made  three  times  as  large  as  the  exhaust  valve  area  of  the  ordinary 
counterflow  engine.  If  now  the  remaining  cross-sections  have  sufficient  area  to 
harmonize  with  these  large  exhaust  port  areas,  and  the  length  of  the  passages  is 
kept  down  to  the  minimum,  then  complete  pressure  equalization  will  result.  This 
is  proved  by  experience  as  well  as  theory  (See  end  of  this  chapter). 

In  Figs.  2  and  3  are  shown  a  longitudinal  and  a  cross  section  of  a  una-flow 
cylinder  where  the  exhaust  belt  connects  over  its  full  width  to  the  jet  condenser 
placed  immediately  below  it.  The  injection  water  enters  by  means  of  a  perforated 
tube  placed  horizontally  across  the  condenser.  As  may  be  seen  from  these  illustra- 
tions, the  exhaust  passages  are  extremely  short  and  wide  so  that  there  is  practically 
no  resistance. 


Fig.  2. 


Fig.  3. 


Figs.  4  and  5  show  the  application  of  a  jet  condenser  of  the  Westinghouse- 
Leblanc  type  to  a  una-flow  cylinder.  This  condenser  and  a  similar  one  developed 
by  the  A.E.  G.  are  based  upon  a  principle  which  formed  the  substance  of  a  patent 
issued  to  the  author.  The  condenser  body  in  this  case  forms  the  support  for  the 
engine  cylinder.  As  in  Figs.  2  and  3,  this  gives  a  very  short  connection  and  large 
transfer  area,  thus  insuring  equalization  of  pressure  between  the  cylinder  and 
condenser. 

On  account  of  this  complete  equalization,  the  compression  begins  at  the  lowest 
possible  pressure  with  the  result  of  a  considerable  gain  of  diagram  area,  a  corre- 
sponding reduction  of  clearance  volume  and  clearance  surfaces,  as  well  as  increased 
thermal  efficiency  (Fig.  1).  The  short  duration  of  the  exhaust  period  due  to  the 
piston-controlled  exhaust  correspondingly  reduces  the  cooling  action  of  the  con- 
denser upon  the  interior  of  the  cylinder.  As  soon  as  the  exhaust  ports  are  covered 
on  the  return  stroke,  the  connection  with  the  condenser  is  cut  off,  any  further 
cooling  is  prevented  and  the  heating  effect  of  the  steam  jacket  at  the  cylinder 
end  comes  into  full  play  without  any  adverse  influence  due  to  the  exhaust. 

It  is  fundamentally  wrong  to  interpose  oil  separators,  change-over  valves, 
feed  water  heaters  or  elbows  in  the  connection  between  engine  cylinder  and  con- 


70 

denser.  Such  accessories  cause  very  large  resistances  to  the  flow  of  steam  and  should 
be  avoided  unless  their  use  is  rendered  necessary  by  other  important  considera- 
tions. 

The  atmospheric  exhaust  pipe  should  be  connected  to  the  condenser  body. 
If  the  connection  between  the  condenser  and  air  pump  is  shut  off,  the  former 


Fig.  4. 


Fig.  5. 


then  acts  as  a  kind  of  exhaust  muffler  or  silencer  (See  Fig.  2).  This  silencer  action 
should  be  assisted  not  only  by  the  volume  of  the  condenser  but  also  by  a  change 
of  direction  of  the  steam  flow.  If  no  such  provision  is  made,  the  loud  exhaust 
will  be  very  objectionable,  as  is  shown  by  experience. 


»<.  <^  ^  CQ  PdJP  uoj  sy/^i/o 


cr-p 


71 


4.  Losses  due  to  Friction.    (Mechanical  Efficiency.) 
Dimensioning  of  Driving  Parts. 

Very  complete  data  are  available  for  the  dimensioning  of  driving  parts  of 
stationary  una-flow  engines.  From  these  data  have  been  compiled  the  curves 
shown  in  Fig.  1,  which  apply  to  steam  pressures  of  from  10  to  12  at.  gage,  and 
condensing  operation. 

In  Fig.  1  may  be  seen  the  weight  of  the  reciprocating  parts  including  two- 
thirds  of  the  weight  of  the  connecting  rod,  plotted  against  cylinder  diameter. 
The  average  values  may  be  represented  by  a  curve  according  to  the  equation 

r      __  (Cylinder  dia.  in  cm)2-5 
&VT  -26- 

The  weight  of  the  piston  is  about  3/10,  that  of  the  connecting  rod  2/7  °f  the  total 
weight  of  the  reciprocating  parts. 

Since  the  ratio  of  stroke  to  cylinder  bore  is  the  determining  factor  for  the 
reciprocating  weights,  the  ratio  of  stroke  to  cylinder  bore  is  also  shown  in  this 
chart.  It  will  be  observed  that  small  engines  have  a  proportionally  long  stroke, 
while  large  engines  have  a  proportionally  shorter  stroke.  Since  the  average  buyer 
of  engines  generally  has  a  prejudice  against  what  may  be  called  high  speed  in 
the  sense  of  high  number  of  revolutions  per  minute,  regardless  of  piston  speed, 
small  engines  are  therefore  built  with  a  comparatively  long  stroke.  For  large 
engines  a  high  number  of  revolutions  is  usually  demanded,  and  since  the  majority 
of  builders  have  a  similar  dislike  for  high  piston  speeds,  an  engine  of  short  stroke 
is  the  result.  It  seems  strange,  however,  that  the  type  of  frame  used  does  not 
appear  to  have  any  bearing  whatever  upon  the  bore  and  stroke  ratio  although 
some  designers  are  inclined  to  make  side  crank  engines  with  long,  and  center 
crank  engines  with  short  strokes  (See  Fig.  1). 

The  piston  speed  based  on  cylinder  diameter  shows  a  more  rapid  increase 
for  the  smaller  sizes  than  for  the  larger  ones. 

The  maximum  continuous  load  rating  of  una-flow  engines  usually  corre- 
sponds to  a  mean  effective  pressure  of  about  4.5  kg/sqcm.  If  the  mechanical  effi- 
ciency for  this  load  is  assumed  to  be  0.94,  then  the  figures  for  the  HP  output 
obtained  agree  closely  with  those  given  by  the  makers.  The  normal  rating  usually 
corresponds  to  a  mean  effective  pressure  of  3  kg/sqcm  based  on  brake  HP.  This 
figure  is  evidently  a  compromise  between  high  economy  and  low  initial  cost. 

The  lowest  curve  in  Fig.  1  represents  weight  of  reciprocating  parts  divided 
by  rated  brake  HP.  The  values  show  small  variation  (3.45  to  4.15  kg/BHP)  but 
increase  gradually  with  the  cylinder  bore. 

The  curve  between  3.6  and  5.6  kg/sqcm  represents  inertia  of  reciprocating 
parts  for  an  infinite  length  of  connecting  rod. 


72 


Fig.  2  gives  first  an  indicator  card  having  an  MEP  of  4.3  kg/sqcm  corre- 
sponding to  the  maximum  continuous  load.  From  this  card  are  developed  three 
net  pressure  diagrams  containing  inertia  curves  plotted  for  a  length  of  connecting 
rod  equal  to  five  times  the  crank  radius,  for  three  different  cylinder  bores  of  500> 

Net  pressure  and  Inertia  force  Curves. 


MEP  =  4,3  kg/ cm 


10- 


Cylinder  dia.  500  mm 


Crank  pin 

Main  bearing,  for 

50%  balancing 


Cylinder  dia.  900  mm 


Piston  rod 

Crosshead  pin 


Fig.  2. 

900  and  1300  mm.  For  any  piston  position  the  vertical  distance  between  inertia 
curve  and  net  pressure  line  represents  the  load  upon  that  driving  part  to  which 
the  inertia  curve  refers.  The  dashed  and  dotted  lines  apply  to  the  load  on  the 
piston  rod,  the  dashed  lines  to  the  crosshead  pin,  and  the  full  lines  to  the  crank 
pin.  The  dotted  curves  in  the  same  way  give  the  load  on  the  main  bearings,  50% 
of  the  weight  of  the  reciprocating  parts  being  balanced.  For  center  crank  shafts 
two  equally  loaded  main  bearings  of  equal  size  are  assumed,  while  for  side  crank 


73 

shafts  the  main  bearing  loads  have  been  increased  by  about  20%  on  account  of 
the  overhang.  In  regard  to  bearing  load,  apart  from  impact,  it  would  be  more 
advantageous  if  the  inertia  forces  for  the  smaller  engines  would  correspond  to 
those  of  the  larger  size  in  the  last  diagram  of  Fig.  2;  and  this  could  be  accom- 
plished by  increasing  the  speed,  for  an  engine  of  500  mm  cylinder  bore,  from 
162  r.  p.  m.  to  say  188  r.  p.  m. 

The  proportions  of  piston  rod  and  tail  rod  as  well  as  diameters  of  the  diffe- 
rent bearings  are  plotted  as  functions  of  the  cylinder  bore.  It  will  be  noted  that 
the  ratio  of  piston  rod  diameter  to  cylinder  bore  is  slightly  less  for  engines  of 
large  size,  by  reason  of  the  proportionally  shorter  stroke  of  the  latter.  The  dis- 
tance from  rear  end  of  piston  to  center  of  crosshead  pin  is  usually  about  3.33  times 
the  stroke.  The  factor  of  safety  against  buckling  of  the  piston  rod,  based  on  the 
loads  represented  in  the  diagram  of  Fig.  2,  is  10  for  small  engines  and  9  for  larger 
engines. 

Average  ratios. 

Crosshead  pin  diameter  to  cylinder  bore 0.265 

Side  crank,  crank  pin  diameter  to  cylinder  bore 0.33 

Side  crank,  main  bearing  diameter  to  cylinder  bore  .    .    .  0.5 

Center  crank,  crank  pin  diameter  to  cylinder  bore     .    .    .  0.425 

Center  crank,  main  bearing  diameter  to  cylinder  bore   .    .  0.425 

Length  of  crosshead  pin  to  its  diameter 1.3 — 1.6 

Side  crank,  crank  pin  length  to  its  diameter    .    ...    .    .1.0- — 1.2 

Center  crank,  crank  pin  length  to  its  diameter 0.9 — 1.0 

Side  crank,  main  bearing  length  to  its  diameter 1.3 — 1.8 

The  crank  pin  diameter  of  center  crank  shafts  will  be  found  only  in  rare  cases 
to  be  larger  than  the  diameter  of  the  main  bearing,  and  the  main  bearing  at  the 
flywheel  side  longer  than  the  opposite  main  bearing. 

F 
Fig.  1  further  shows  the  ratios  of  piston  area  to  bearing  areas  — — —  which 

(I '  a) 

give  the  following  averages:  Crosshead  pin  8,  side  crank,  crank  pin  7. 7,  center  crank, 
crank  pin  4.25,  side  crank,  main  bearing  1.95,  and  each  main  bearing  of  center 
crank  shafts  1.5. 

Combining  these  data  with  the  specific  loading  taken  from  the  diagrams  in 
Fig.  2,  the  resultant  bearing  pressures  were  calculated  and  are  shown  at  the  top 
of  Fig.  1.  These  values  refer  to  maximum  continuous  load  and  smallest  dead 
center  inertia,  horizontal  forces  only  being  considered.  Strictly  speaking,  the 
additional  forces  due  to  flywheel  weight,  belt  pull,  etc.  should  be  combined  with 
the  horizontal  forces,  but  this  would  not  materially  alter  the  results.  It  will  be  seen 
that  crank  pin  and  main  bearings  of  side  crank  shafts  sustain  about  50%  higher 
loading  than  the  corresponding  bearings  of  center  crank  shafts.  The  highest  bearing 
pressures  given  in  Fig.  1  are  undoubtedly  permissible  in  case  of  force  feed  lubri- 
cation. 

On  side  crank  shafts,  excessively  large  leverages,  i.  e.  distances  from  the  con- 
necting rod  center  to  the  main  bearing  center,  cause  increased  bending,  higher 


74 


bearing  pressure  and  on  account  of  the  deflection  of  the  crank  shaft,  increased 
pressure  at  the  inside  edge  of  the  main  bearing.  This  tendency  can  be  reduced 
by  shortening  the  leverage  or  the  use  of  self  aligning  bearings,  or  both  (Fig.  3). 
Since,  there  are  no  secondary  forces  acting  on  the  connecting  rod  in  the  horizontal 
plane,  the  factor  of  safety  against  buckling  in  this  plane  need  not  be  more  than  5, 
as  against  a  factor  of  9  or  10  in  the  vertical  plane.  This  condition  can  be  easily 
met  by  flattening  the  otherwise  circular  rod  section.  The  crank  hub  should  be 

placed  as  close  as  possible 
to  the  connecting  rod  and 
should  not  be  wider  than 
0,65  times  the  shaft  dia- 
meter for  large  engines 
and  0.75  for  small  engines. 
On  some  large  Belgian 
engines  this  figure  is  even 
cut  down  to  0.6.  The 
projecting  part  of  the 
crank  pin  bearing  cor- 
responds to  the  projecting 
part  of  the  crank  hub. 
The  rear  side  of  the  crank 
in  this  case  becomes  flat. 
The  crank  is  frequently 
pressed  in  place  on  the 
shaft,  although  a  shrink 
fit  is  preferable  by  reason 
of  the  lesser  chance  of 
damage  to  the  structure 
of  the  material.  A  key, 
although  frequently  used, 
is  unnecessary.  The  same 
applies  to  the  connection 
of  the  crank  pin  to  the 
crank  arm,  and  the  former 
should  be  ground  after 
assembly. 

Fi8-  3-  UsssL^J  A   still   shorter   over- 

hang may  be  obtained  by 
casting  crank,  crank  pin 
and  crank  shaft  in  one 

piece  of  cast  steel.  <  By  designing  the  crank  in  the  form  of  a  disc,  as  shown  in 
Fig.  4,  an  especially  ,large  reduction  in  the  overhang  may  be  obtained,  with  a 
corresponding  decrease  in  the  shaft  diameter.  The  present  state  of  foundry  prac- 
tice allows  such  a  construction  to  be  used  without  anxiety. 

Constructional  data  of  a  stationary  una-flow  engine  built  by  Sulzer  Bros, 
and  installed  in  a  cotton  spinning  mill  at  Grefeld  are  as  follows:  Cylinder  bore 


|  23000  Kc) 


75 

1100  mm,  stroke  1200  mm,  speed  110  r.  p.  m.,  steam  pressure  12  at.  gage,  steam 

temperature  320°.    The  engine  is  of  the  center  crank  type. 
Main  bearings,  475  mm  dia.  by  650  mm  long 
Crank  pin,  475  mm  dia.  by  380  mm  long 
Crosshead  pin,  300  mm  dia.  by  430  mm  long 
Piston  rod  220  mm  diameter,  tail  rod  170  mm  diameter 
Weight  of  piston  2125  kg,  weight  of  piston  rod  1500  kg 
Weight  of  crosshead  1532  kg,  2/3  connecting  rod  1780  kg 
50%  of  the  reciprocating  parts  are  balanced 


by  counterweights  fastened  to  the  crank 
cheeks.j                    \ 
The  friction  HP  of  this  engine  at  110  r.  p.  m. 
with  a  smooth  flywheel  was  113,6.    The  correspon- 
ding mechanical  efficiency  for  a  rated  load  of  1700  1  HP 
is   therefore   0.933,    a    figure    which   disproves   the 
opinion   frequently  advanced   that    the   mechanical 
efficiency  of  una-flow  engines  is  low.  The  assumption 
that  the  engine  friction  must  be  nearly  independent 
of  the  HP  output  is  based  on  the  fact  that  the  load 
on  the  driving  parts  is  practically 
the  same   for  idling  as  it  is  for                   K 
the  rated  HP     Starting  with  the                   '-  

-j 

^ 

^ 

^ 

ss/s 

1 

\ 

^ 
^r 

1: 

^ 

1 
1 

1 

^ 
ti 

•OOx 

SS> 
^s 

xxX 

^ 
^ 

^N: 
^t 

mi 

- 

**— 

engine  idling,    and   gradually  in- 
creasing  the  output,    the    gross         \ 

^ 

load  on   the   driving   parts   first        A 
decreases  slightly,  at  rated  output        ^ 
reaches   the    same    value    as  for        \^ 
idling,  and  then  becomes  some- 
what greater   for   larger  output. 
The  engine    at  Crefeld,    for  instance,    has 
center   inertia   load    of   approximately   6  ] 
a  corresponding  mean  pressure   of   the  ine 
gram  of  3  kg/sqcm,  and  a  useful  rated  me; 
tive   pressure   of  3   kg/sqcm.      It  follows  1 
engine   friction   for  idling   and   rated  outp 
be  the  same.     Further,   the  weight   of  the 
flywheel,  crank  shaft  and  the  rest   of  the 
parts  as  well  as  the  centrifugal  force  oftheco 
rod  end,  crank  and  counterweights  are  res 
for  a  constant  portion  of  the  total  friction 
The   above  will  be   further  illustrated 
following  test  results  obtained  by  Sulzer  B 

\ 

^  /  ^^  /  /'s 

a   dead 
cg/sqcm, 
rtia  dia- 
m  effec- 
,hat  the 
ut  must 
piston, 
driving 
rmecting 
ponsible 

by  the 
ros. 

Fi 

/s 

I 

->v 

1 

j 

g- 

4. 

A  una-flow  engine  of  700  mm  cylinder  bore  and  900  mm  stroke,  having  a  rope 
flywheel  of  4000  mm  diameter,  gave  the  following  friction  at  different  speeds  (without 
ropes).  133  112  100  85  68  r.  p.  m. 

66  53  46  38  28  I  HP  friction. 


76 

The  corresponding  rated  load  would  be 

820  690  615  520  420  I  HP, 

so  that  the  friction  HP  in  %  would  be 

8  7.7  7.5  7.3  6.7%. 

Another    engine    of    600   mm    cylinder  bore    and    725   mm    stroke,    fitted 

with  a  rope  flywheel  of  2400  mm  diameter,  with  14  grooves,  gave  the  following 

results:                         150                100  50                34  r.  p.  m. 

41                  22  8.5             4.5    I  HP  friction. 

The  corresponding  rated  load  in  this  case  would  be 

475  317  158  108  I  HP, 

so  that  the  friction  HP  in   %  is 

8.65  7  5.4  4.15%. 

In  reducing  the  engine  speed  from  150  to  50  r.  p.  m.  the  friction  HP  should 
diminish  from  41  HP  to  1/9  of  the  same,  or  4.5  HP.  It  actually  was  8.5  HP,  which 
reflects  the  influence  of  the  weight  of  the  driving  parts.  The  weight  and  inertia 
of  the  latter  have  an  equalizing  effect,  so  that  for  constant  speed  the  friction  HP 
remains  nearly  constant  independently  of  the  instantaneous  output. 

Dimensions  of  Driving  Parts  of  Other  Una-Flow  Engines- 
(Sulzer  Bros,  design.) 

Steam  pressure  at  admission  valves  12  at.  gage.  All  bearings  have  force  feed 
lubrication. 

1.  550  mm  cylinder  bore,  650  mm  stroke,  158  r.  p.  m. 

Two  main  bearings 250  mm  dia.    by  360  mm  long 

Crank  pin .    .    , .   .  250    „       „       „    200    „ 

Grosshead  pin 150    „       „       „    220    „ 

Grosshead  shoes 520    „     long    „    300    „     wide 

Piston  rod 110    „     dia. 

Connecting  rod  length   ....  5.5  times  the  crank  radius. 

2.  500  mm  cylinder  bore,  600  mm  stroke,  165  r.  p.  m. 

Two  main  bearings     .....   230  mm  dia.    by  330  mm  long 

Crank  pin 230    „       „       „    180     „ 

Crosshead  pin 140    „       „       „    200    „ 

Crosshead  shoes 480    „     long    „    275    „     wide 

Piston  rod 100    „     dia. 

Connecting  rod  length   ....    5.5  times  the  crank  radius. 

3.  850  mm  cylinder  bore,  1000  mm  stroke,  125  r.  p.  m. 

Two  main  bearings 375  mm  dia.    by  550  mm  long 

Crank  pin 375    „       „       „    310    „        „ 

Crosshead  pin 240    „       „       ,,    350    ,, 

Crosshead  shoes 800    „     long    „    500    „     wide 

Piston  rod 150    „     dia. 

Connecting  rod  length   ....    5.5  times  the  crank  radius. 


77 

Sulzer  Bros,  also  mention  the  fact  that  they  have  found  the  friction  HP  of 
their  una-flow  engines  equal  or  slightly  less  than  the  friction  HP  of  their  tandem 
compound  engines  of  equal  power. 

The  driving  parts  of  a  una-flow  engine  should  naturally  cause  more  friction 
than  the  parts  of  a  tandem  compound  engine  since  the  size,  or  rather  diameter, 
of  the  bearings  of  a  una-flow  engine  is  larger  on  account  of  the  higher  piston  load. 
In  a  una-flow  engine  the  single  piston  carries  live  steam  pressure,  while  the  low 
pressure  piston  of  a  tandem  compound  engine  only  carries  receiver  pressure  and 
the  much  smaller  high  pressure  piston  sustains  the  difference  between  the  live 
steam  and  receiver  pressures.  On  the  other  hand,  the  single  una-flow  cylinder 
with  its  one  piston  and  one  or  two  piston  rod  packings  will  cause  less  friction  than 
the  two  cylinders  with  two  pistons  and  three  or  four  rod  packings  of  the  tandem 
counterflow  engine.  Steam  cylinders  arranged  in  tandem  are  furthermore  subject 
to  misalignment  with  accompanying  binding  of  the  moving  parts,  and  the  friction 


Fig.  5. 

caused  by  such  misalignment  may  be  considerable.  The  piston  system  in  una- 
flow  engines  can  always  be  supported  at  two  points  only,  for  instance  by  means 
of  a  crosshead  and  self-supporting  piston,  or  on  crosshead  and  tail  rod  support 
with  floating  piston.  It  is  assumed  of  course  that  the  metallic  packing  used  is  of 
such  design  as  to  permit  of  lateral  movement  of  the  piston  rod.  Furthermore, 
the  tandem  counterflow  engine  requires  four  times  the  number  of  steam  distri- 
buting elements  as  the  una-flow  engine  (8  valves  against  2  valves),  with  a  corre- 
sponding increase  in  friction.  A  comparison  between  a  una-flow  and  a  cross-com- 
pound engine  will  still  further  emphasize  the  advantages  of  the  una-flow  system. 
According  to  what  is  said  above,  the  lesser  friction  of  the  una-flow  cylinder  must 
outweigh  the  increased  friction  of  the  una-flow  driving  parts,  if  the  experience 
of  Sulzer  Bros,  is  accepted  as  having  general  application. 

That  part  of  the  engine  friction  caused  by  the  piston,  especially  if  self-sup- 
porting, may  be  considerable.  This  friction  may  be  reduced  by  fitting  the  piston 
with  shoes  of  bronze,  babbitt  or  Allan  Metal.  The  friction  is  least  with  a  floating 
piston,  i.  e.  a  piston  supported  by  its  rod,  despite  the  additional  friction  of  the 
second  stuffing  box  and  tail  rod  support. 


78     • 

The  una-flow  piston  should  be  made  as  light  as  possible,  and  this  may  be 
accomplished  by  constructing  it  in  two  parts  of  cast  steel  (See  Fig.  5),  thereby 
reducing  friction,  inertia  and  impact. 

Next  to  the  piston,  the  main  bearing,  crank  pin  and  crosshead  pin  contribute 
the  largest  share  to  the  total  friction.  The  friction  of  high  grade  metallic  packings 
is  extremely  small,  as  is  also  the  friction  of  the  poppet  valves  and  their  gear.  This 
applies  especially  to  una-flow  engines  with  only  two  valves  and  one  packing. 

The  friction  of  the  driving  parts  can  be  considerably  reduced  by  a  proper 
oiling  system,  especially  by  means  of  force  feed  lubrication.  The  latter  type 
also  reduces  impact.  The  provision  of  force  feed  lubrication  for  the  condenser 
pump  driving  parts  and  the  use  of  a  housing  around  the  flywheel  are  further  mea- 
sures in  the  right  direction. 

Short  stroke  engines  with  bearings  of  large  diameter  naturally  have  a  higher 
friction  loss  than  long  stroke  engines. 

In  engines  having  steam  jackets  on  the  cylinder  barrel  the  friction  is  usually 
greater  if  the  jackets  are  shut  off.  In  the  same  way  a  new  engine  while  being  run 
in  will  have  more  friction  than  later,  and  the  friction  of  an  engine  immediately 
after  starting  will  be  larger  than  when  in  regular  operation,  especially  if  it  has 
not  been  warmed  up  previously. 

Taking  it  altogether  one  may  say  that  the  una-flow  engine  has  a  slightly  better 
mechanical  efficiency  than  the  ordinary  tandem  compound  engine. 


79 


5.  Losses  due  to  Leakage. 

Valves,  Pistons,  Piston  Rod  Packings. 

Tight  steam  distributing  organs  are  a  rare  exception.  Slide  valves  are  gene- 
rally considered  to  be  tighter  than  piston  valves,  this  being  the  reason  for  the 
practice  of  many  concerns  to  use  piston  valves  for  the  high  pressure  and  slide 
valves  for  the  low  pressure  cylinders,  a  practice  which  is  also  supported  by  pres- 
sure and  temperature  considerations. 

Corliss  valves  are  fairly  tight,  but  they  are  far  from  being  absolutely  tight. 

Well  made  piston  valves  fitted  with  snap  rings  may  be  considered  fairly  tight. 
Piston  valves  without  rings  should  only  be  used  in  small  sizes  for  saturated  steam 
and  must  be  made  a  good  fit.  Larger  piston  valves  for  use  with  superheated  steam 
should  always  be  equipped  with  snap  rings  on 
account  of  the  necessary  clearance  required  for 
expansion.  Even  then  a  certain  amount  of 
leakage  .must  be  expected  in  those  positions  in 
which  none  or  one  ring  only  is  active,  in 
addition  to  the  constant  amount  of  leakage 
past  the  ring  joints.  In  case  of  highly 
superheated  steam,  carbonized  oil  may  be 
the  cause  of  increasing  leakage. 

Double  beat  valves  are  usually  leaky.  The 
leakiness  increases  with  the  amount  of  balance, 
the  pressure  and  the  temperature.  With  all 
types  of  valves  leakiness  will  be  enhanced  with 
increasing  superheat,  on  account  of  warping 
and  the  increasing  fluidity  of  the  steam. 

The  body  of  double-beat  valves  as  shown 
in  Figs,  1,  2  and  3  will  sustain  a  heavy  load  in 
the  direction  of  the  axis  during  the  expansion 
and  exhaust  periods.  The  corresponding 
deflection  will  cause  the  lower  face  of  the  valve 
to  leave  its  seat  and  leak.  The  radial  forces 
can  be  neglected  if  the  seats  are  made  flat. 

In  the  same  way,  if  the  temperature  of  the  valve  is  higher  than  the  tempe- 
rature of  the  material  forming  its  housing  and  seat,  then  the  valve  body  will 
expand  more  than  the  latter,  and  the  upper  valve  face  will  lose  contact  and  start 
to  leak.  The  above  temperature  difference  may  have  several  causes.  In  a  valve 
design  as  shown  in  Fig.  1,  in  which  valve  and  seat  have  the  same  height  and  the 
same  thickness  of  material,  equal  expansion,  in  the  most  favorable  case,  will 


Fig.  1. 


Fig.  2. 


80 


occur  only  if  the  material  of  both  parts  has  the  same  coefficient  of  expansion. 
This  can  be  realized  by  due  attention  to  the  work  of  the  foundry.  Both  parts  are 
exposed  to  live  steam  temperature  on  one  side  and  to  the  varying  temperature 
of  the  cylinder  steam  on  the  other. 

Conditions  are  not  so  good  in  the  design  shown  in  Fig.  2,  in  which  a  valve 
cage  of  the  ordinary  type  is  used.  Valve  and  seat  are  of  different  thicknesses  and 
therefore  expand  unequally,  especially  during  the  first  period  after  starting.  Fur- 
thermore, the  two  parts  will  have  different  temperatures  during  operation,  since 
the  valve  is  exposed  on  one  side  to  live  steam,  while  the  cage  is  entirely  surrounded 
by  cylinder  steam. 

The  most  unfavorable  design  is  the  one  shown  in  Fig.  1,  page  4,  which 
has  the  valve  seats  cast  in  one  piece  with  the  cylinder  head.  The  difference  in 
expansion  between  valve  and  seat,  especially  just  after  starting  up,  must 
be  considerable.  The  coefficients  of  expansion  and  the  mean  tempera- 
ture of  both  parts  will  certainly  be  different  and  the  corresponding 
leakage  will  therefore  be  considerable. 


Fig.  3. 


Fig.  4. 


The  valve  will  leak  the  more,  the  higher  it  is.  The  first  rule  in  poppet  valve 
design  is  therefore  to  make  the  valve  as  low  as  possible.  For  this  reason,  valve 
gears  should  be  avoided  which  at  late  cut-offs  give  unnecessarily  large  valve  lifts 
and  therefore  require  high  valves,  unless  a  restriction  of  the  upper  passage  is  not 
objectionable.  It  is  further  advisable  in  cases  where  valve  cages  or  similar  con- 
structions are  objected  to  on  the  ground  of  the  number  of  tight  joints  required,  to 
use  a  kind  of  saucer  to  form  the  lower  valve  seat,  thus  at  the  same  time  reducing 
the  height  of  the  valve.  (See  Fig.  3.)  In  this  design  the  vertical  forces  as  well 
as  the  deflection  of  the  valve  body  are  reduced  on  account  of  its  small  height 
and  the  small  radial  width  of  the  valve  ring.  By  grinding  in  at  the  operating 
temperature,  a  close  approximation  to  complete  tightness  for  one  particular  pres- 
sure and  temperature  will  be  obtained.  During  the  first  period  after  starting  and 
at  any  other  pressure  and  temperature  the  valve  will  leak.  Perfect  tightness  under 


81 

all  conditions  can  be  accomplished  with  a  resilient  valve  such  as  is  shown  in  Fig.  4. 
The  lower  valve  face  is  smaller  in  diameter  than  the  upper  one.  This  difference 
produces  a  force  in  addition  to  the  spring  load,  which  tends  to  press  the  lower 
rigid  face  as  well  as  the  resilient  upper  face  against  the  corresponding  seats.  The 
upper  resilient  face  also  takes  care  of  unequal  expansion.  In  order  to  make  the 
valve  low  and  to  reduce  its  deflection,  a  saucer  forming  the  lower  valve  seat  may 
be  used  to  advantage.  Dirt  in  the  steam,  however,  may  cause  leakage  even  with 
this  type  of  valve. 

Theory  of  the  Resilient  Valve. 

The  following  forces  act  on  the  valve: 
Downwards: 

1.  The  spring  pressure  minus  the  steam  pressure  on  the  stem, 


where  pa  is  the  absolute  pressure  in  the  valve  chest. 
2.  The  pressure  on  the  upper  side  of  the  resilient  annular  ring, 


where  pt  is  the  absolute  pressure  in  the  cylinder. 
Upwards  : 

3.  The  pressure  on  the  lower  side  of  the  lower  annulus 

(r*  —  Q*)n(pa~Pi). 

4.  The  upward  reaction  of  the  upper  valve  seat  Wv 

5.  The  upward  reaction  of  the  lower  valve  seat  Wz. 
The  horizontal  forces  balance  each  other. 

The  sum  of  the  vertical  forces  must  be  zero,  so  that 

P  +  (&  —  Qz)n(Pa  —  Pi)-(r*  —  Q*)n(pa-Pi)  —  W,—  W2  =  0       .     (1) 

If  the  radii,  the  steam  pressures  and  the  spring  pressure  are  known,  only 
the  bearing  reactions  W1  and  Wz  remain  unknown. 

W1  may  be  found  from  the  condition  that  the  deflection  f±  of  the  resilient 
ring  due  to  the  steam  pressure  (measured  at  the  middle  of  its  face),  must  be  equal 
to  the  deflection  /2  caused  by  the  seat  reaction  W^  provided  the  lower  valve  face 
is  to  remain  on  its  seat.  Therefore  fl  =  /2. 

Imagining  a  piece  cut  out  radially  from  the  valve  with  the  angle  dcp  at  the 
center,  then  the  steam  pressure  acting  on  this  element  of  the  resilient  surface  is 


and  the  corresponding  deflection  is 

*»>•(*-<>)  (p.  -p0  ....    (2) 


where  E  =  the  modulus  of  elasticity  and   J  =  the  moment  of  inertia  (approxi- 
mately  constant)   of  the  cross-section   at  right  angles   to  the  plane   of  bending. 

Stumpf,  The  una-flow  steam  engine.  6 


82 

The  seat  reaction  on  the  radial  element  under  consideration  is  W1  •  H^-.  and 

2n 

the  corresponding  deflection  is 

^ =   Wl'  In'    3E-J       '     '     ' (3) 

Then  on  equating  /x  =  /2 

Wl  =  -Q-  n  (R  -\-  Q)  (R  —  Q)  (pa  —  p^ (4) 

From  equation  (1) 

Substituting  in  (5)  the  value  of  W±  as  found  in  (4),  then 

In  the  limit,  when  the  valve  just  rests  on  the  lower  seat,  W2  =  0.  Neglecting 
P,  the  excess  of  the  spring  pressure  over  the  steam  pressure  on  the  valve  stem, 
the  highest  permissible  value  for  Q  may  be  obtained  from  the  following  equation: 

"8"       '~r   +  ~s^'-  <7) 

Denoting  R  by  Q  -f-  a,  and  r  by  Q  -f-  6,  equation  (7)  becomes 


10 


or  since  a-\-1q  and  6  -f-  2p  are  approximately  equal, 


If  the  valve  expands  by  the  amount  A  I  in  excess  of  the  casing,  in  consequence 
of  unequal  temperatures  and  coefficients  of  expansion,  the  resilient  ring  must 
deflect  by  the  same  amount  in  order  to  remain  steam  tight.  In  this  case  /,  —  /2 

is  not  zero  but  is  equal  to  41  _  ./        /  /cn 

//t  —  /i  —  /a      .........     •     (v) 


The  seat  reaction  W-^  for  any  given  A  I  is  obtained  by  inserting  the  values 
for  /j  (from  equation  2)  and  /2  (from  equation  3)  in  equation  9;  hence 


For  the  valve  to  be  tight,  W1  must  be  positive,  for  which  purpose  a  pressure 
difference  (pa  —  Pi)  is  necessary,  as  may  be  seen  from  equation  (10).  The  mini- 
mum pressure  difference  required  for  tightness  may  be  obtained  from  equation 
(10)  by  making  W^  —  0,  i.  e. 

d 


83 


From  this  equation  it  will  be  seen  that  the  pressure  difference  is  proportional 
to  Al  and  consequently  to  1.  In  order  to  keep  (pa — />»)  small,  the  valve  must 
be  low.  The  dimension  d  is  determined  by  the  strength  of  the  material,  and  the 
values  of  R  and  Q  by  the  desired  steam  velocity. 

To  find  the  bending  stresses  of  the  resilient  ring  during  operating  condi- 
tions; let 

^  =  the  temperature  of  the  valve, 

Q!  =  the  coefficient  of  expansion  of  the  valve  material, 
t2  =  the  temperature  of  the  surrounding  casing, 
a2  =  the  coefficient  of  expansion  of  the  casing, 
I  =  the  distance  between  valve  seats  at  normal  temperature  t0. 

Then  Al  =  l\a  (t  t) a  (t  Ml  (12) 

Substituting  this  value  in  equation  (11),  the  pressure  difference  (pa  —  p^ 
will  be  found  at  which  the  valve  will  commence  to  be  tight.  At  all  greater  pressure 
differences  the  valve  will  be  perfectly  tight. 

For  example,  a  valve  is  taken  in  which  /  =  30  mm,  R  =  125  mm,  Q  =  104  mm, 
d  —  3  mm,  pa  =  12  at,  and  pt  =  0  at. 

Assume  also  that    t0  =15° 
*!  -  300° 
%  =  0.000012 
a2  =  0.000011. 

The  following  table  gives  the  relative  expansions  Al  of  this  valve  for  certain 
temperature  differences  (t±  —  Z2),  calculated  according  to  equation  12.  In  this 
table  are  also  found  the  stresses  K^  and  the  pressure  differences  (pa  —  p^  neces- 
sary for  tightness  calculated  by  means  of  equation  11. 


tl        tz  — 

0° 

50° 

100° 

150° 

200° 

Jl  = 

in  mm 

0.009 

0.0255 

0.042 

0.0585 

0.075 

Pa  ~  Pi  = 
in  at  abs. 

1.35 

3.8 

6.3 

8.75 

11.25 

K»  = 

230 

645 

1070 

1490 

1920 

From  these  figures  it  will  be  seen  that  many  prevalent  valve  constructions 
are  far  from  being  steam  tight.  The  excessive  height  of  many  valves,  neglecting 
other  constructional  errors,  makes  steam  tightness  absolutely  impossible.  Even 
in  a  valve  only  30  mm  high,  with  a  temperature  difference  between  valve  and 
seat  of  200°  G.,  a  steam  pressure  of  11.25  kg/sqcm  is  necessary  in  order  to  obtain 
steam  tightness;  i.  e.  at  all  lower  pressures  there  will  be  leakage. 

6* 


84 

Calculations    for    a    Resilient   Valve    of    160  mm    mean    diameter,    for    a 
condensing  engine   working  with   steam  of  9  at.  abs.   (Fig.  5). 

The  following  two  condi- 
tions are  assumed  for  the  cal- 
culations of  Wl  and  W2: 

1.  The  actual  closure  oc- 
curs at  the  mean  circum- 
ference in  both  the  top 
and  bottom  seats  (Radii 
Rm  and  rm). 

2.  In  the  most  unfavorable 
case   closure   occurs   at 
the   inner   edge   of  the 
upper  seat  (Ri)  and  the 
outer  edge  of  the  lower 
seat  (ra). 

In  the  latter  case  W2  on 
the  lower  valve  seat  is  to  be 
zero. 


In  Fig.  5, 


Fig.  5. 

=  81.5  mm 

=  76  mm 


Ri  =  80  mm 
ra  —  77.5  mm 

6*71 


Q  =  67.5  mm 
d  =  17  mm. 


The  steam  pressure  on  the  valve  stem  is  — —  •  (pa  —  1)  =18  kg. 

Wj,  according  to  equation  (4)  above,  is 

221  kg,      for  Rm  and  rm 
195.5  kg,  for  Rt    and  ra. 

W2  =  0  in  case  2.    The  spring  pressure  F  is  then  calculated   from  equa- 


tion (5), 


(Pa-1) 


(Pa  -  Pi)  (R?  n-r*.  n] 


Wl 


=  18  +  195.5  —  107  =  165.5  kg. 
With  this  spring  pressure  in  case  (1) 

W2  =  P  +(pa  —  Pi)  (Rm*  n  —  rm2  n 
=  88.5  -f  245  —  221  =  112.5  kg. 

The  thickness  d  of  the  resilient  ring  is  calculated  for  the  case  in  which  the 
valve  is  opened  at  the  greatest  pressure  difference  pa  —  p,,  the  valve  and  seat 
being  assumed  to  have  the  same  temperature. 

Again  taking  a  radial  element  d<p  of  the  valve  (Fig.  4),  then 


and 


. 

(Pa - 


85 

If  d  =  2  mm,  then  the  bending  stress  at  the  radius  Rm  is  Kb  =  1460  kg/sqcm. 

Such  a  high  stress  can  only  occur  when  the  valve  is  opened  against  the  full 
pressure  difference. 

If  the  difference  in  seat  diameters  is  made  less  than  is  required  by  the  above 
theory,  then  the  lower  valve  face  will  lift  off  its  seat  and  leakage  will  result.  It 
is  therefore  always  advisable  to  make  a  second  calculation  with  the  object  of 
finding  W2  under  assumptions  similar  to  case  2  (See  above).  For  this  worst  case 
the  numerical  value  of  W%  should  be  positive. 

The  valve  seats  should  always  be  made  flat  since  they  can  then  accommo- 
date unequal  radial  expansion  and  give  the  maximum  opening  for  the  valve  lift. 
The  latter  is  always  small  in  una-flow  engines  on  account  of  the  early  cut-offs 
and  the  fact  that  the  permissible  lead  is  very  small. 

Experience  shows  that  the  upper  seat  wears  more  than  the  lower,  especially 
if  the  seats  are  narrow.  The  upper  seat  should  therefore  be  made  wider  and  the 
edge  of  the  resilient  ring  reinforced  by  a  rim  or  bead.  The  resilient  ring  should 
not  be  made  too  weak. 

The  piston  of  a  una-flow  engine  is  placed  between  live  steam  space  and  ex- 
haust. Live  steam  space,  inlet  valve,  piston  and  exhaust  are  in  series,  while  in  the 
ordinary  counterflow  engine  the  piston  is  in  parallel  with  the  inlet  and  exhaust 
valves.  This  series  arrangement  is  one  of  the  important  features  of  una-flow  engines. 
The  leakage  of  the  steam  valve  of  a  counterflow  engine  is  partly  balanced  by  the 
leakage  of  the  exhaust  valve  so  that  only  the  difference  of  these  leakages  affects 
the  cylinder;  the  leakage  of  the  inlet  valve  of  the  una-flow  engine  on  the  other 
hand  has  no  such  compensating  means  since  there  is  no  exhaust  valve,  in  addi- 
tion to  this,  the  counterflow  exhaust  valve  usually  remains  open  during  the  greater 
part  of  the  return  stroke  so  that  the  leakage  of  the  inlet  valve  only  affects  a  short 
part  of  the  latter.  In  the  una-flow  engine,  however,  the  leakage  of  the  steam 
valve  continues  during  almost  the  whole  compression  stroke,  and  this  has  to  be 
taken  into  account  by  properly  proportioning  the  clearance  volume,  in  order  that 
the  compression  may  not  exceed  the  initial  pressure  and  the  engine  begin  to  be 
noisy.  The  fact  that  leakage  of  the  inlet  valves  of  una-flow  engines  makes  itself 
noticeable,  especially  with  small  clearances,  must  be  considered  an  advantage 
of  this  type  of  engine,  the  effect  of  which,  however,  may  sometimes  become  em- 
barrassing. Gounterflow  engines  show  no  such  indication,  so  that  great  losses  some- 
times pass  unnoticed.  The  above  mentioned  series  arrangement  together  with 
the  short  duration  of  the  exhaust  tends  to  considerably  reduce  the  losses  due  to 
leakage.  These,  however,  may  still  be  large  if  precautions  are  not  adopted  in  the 
shape  of  resilient  or  single-beat  valves. 

A  resilient  valve  may  also  be  made  of  cast  iron  (Figs.  6  and  7).  The  author 
has  designed  several  such  valves  for  use  on  British  una-flow  engines,  which  have 
given  complete  satisfaction. 

The  necessary  amount  of  resiliency  may  also  be  obtained  by  properly  shaping 
the  lower  seat  on  which  the  valve  rests.  (Fig.  8.)  Here  the  seat  at  the  top  of  the 
saucer  is  formed  with  a  resilient  ring  whose  undercut  should  be  such  that  the 
steam  pressure  acting  upon  it,  in  combination  with  the  valve  spring  and  the  steam 
pressure  on  the  valve  stem  give  the  same  reactions  on  both  seats.  This  design 


86 


offers  the  advantage  that  the  valve  can  be  made  of  cast  iron  in  the  ordinary  way 
and  only  the  relatively  simple  saucer  need  be  made  of  steel. 

The  same  principle  is  ex- 
pressed in  Fig.  9,  in  which 
the  resilient  ring  is  a  sepa- 
rate piece  acting  in  the  same 
way  as  in  the  previous  valves. 
Finally,  the  valve  shown 
in  Fig.  10  obtains  the  neces- 
sary resiliency  by  means  of  a 
separate  seat  plate  fitted  with 
snap  rings  and  pushed  up- 
wards by  small  springs.  Holes 
pro  vided  in  the  seat  plate  allow 
the  live  steam  pressure  to  act 
underneath  it ;  and  as  soon  as 
Fig.  6.  a  pressure  difference  exists  be- 

tween valve  chest  and  cylin- 
der, the  resulting  pressure  on  the  small  annular  ring  outside  the  lower  valve  seat  tends 
to  press  the  seat  plate  with  increasing  force  against  the  valve  face.  The  diameter 
of  this  seat  plate  must  be  proportioned  so  that  the  upward  steam  pressure  com- 
bines with  the  spring  force,  the  steam  pressure  upon  the  valve  stem  and  the  steam 
pressure  upon  the  unbalanced  area  of  the  valve  so  as  to  produce  equal  bearing- 
reactions  on  both  seats.  This  last  design  has  the  least  merit,  since  the  seat  plate  rings 
will  never  be  absolutely  steam  tight,  and  the  clearance  volume  is  larger  than  that 
required  by  the  forms  of  valve  previously  described.  Further  the  springs  necessi- 
tated in  this  case  as  also  those  of  Fig.  9  will 
very  soon  weaken  at  high  steam  temperatures. 
All  of  the  above  constructions  are  based 
upon  the  previous  theory  which  requires  a  certain 
difference  in  diameter  of  the  two  valve  seats. 


Single-Beat  Valves. 

If  one  has  studied  the  valve  question  in 
all  its  phases  and  has  become  acquainted  with 
all  of  the  difficulties  of  manufacture  and 
operation,  then  the  final  result  must  be  con- 
sidered unsatisfactory.  The  following  points  Fig.  7. 
should  be  noted. 

1.  If  a  double-beat  valve  is  completely  or  partly  balanced,  it  completely  or 
partly  violates   the   principles   embodied   in   the  single-beat   valve,   according  to 
which  the  closing  pressure  is  proportional  to  the  area  covered  and  to  the  pressure 
difference.     This    violation  explains  most    of   the  difficulties    usually   met    with. 

2.  If  the  double-beat  valve  is  left  unbalanced  by  such  an  amount  as  is  required 
by  the  resilient  valve  theory,  then,  if  it  were  possible  to  provide  sufficient  lift,  a 
single-beat  valve  of  an  area  equal  to  the  unbalanced  area  of  the  double-beat  valve 


87 


would  provide  sufficient  opening  for  the  steam  flow.    Such  a  single-beat  valve 
with  the  corresponding  valve  gear  is  described  in  the  last  chapter. 

If  no  difficulties  have 
been  encountered  with  the 
resilient  valves  in  respect  to 
governing,  the  same  must  hold 
true  of  the  single-beat  valve 
of  equal  area.  This  forms  an 
accord  with  the  internal  com- 
bustion engine  in  which  the 
single-beat  valve  has  been  per- 
sistently adhered  to  for  very 
good  reasons.  The  high  lift 
single-beat  valve  has  a  diameter 
which  is  less  than  half  of  the 
diameter  of  a  double-beat  valve 
of  equal  area  of  flow.  Such 
a  valve  may  be  ground  in 
cold.  It  allows  of  a  conside- 
rable reduction  of  the  clearance  Fig-  8. 
volume  and  harmful  surfaces, 

and  insures  perfect  tightness  at  all  temperatures  and  pressures  even  with  extreme 
variations  of  the  same.    It  may  therefore  be  claimed  that   the  single-beat  valve 

with  a  high  lift  is  the  only 
means  by  which  complete 
safety  against  leakage  can  be 
provided. 


Valve  Bonnets. 

A  valve  bonnet  of  Lentz 
design  is  shown  in  Fig.  11. 
The  connection  between  stem 
and  valve  is  made  by  means 
of  the  valve  stem  head  and1 
two  nuts,  one  of  them  being 
a  lock  nut.  The  stem  screws 
into  the  cast  iron  roller  head 
and  is  secured  by  a  lock  nut. 
Fine  thread  is  advisable  in 
both  places.  The  valve  stem 
guide  is  formed  by  a  cast  iron 

bushing  which  should  be  made  heavy  to  facilitate  manufacture,  and  in  case  of 
superheated  steam  should  be  equipped  with  a  separate  force  feed  lubricating 
connection.  It  is  wrong  to  supply  oil  to  the  steam  chest.  The  flow  of  steam 
through  the  cast  iron  valve,  on  account  of  the  slight  unavoidable  obliquity 
of  the  ribs,  produces  a  turning  moment  which  must  be  resisted  by  properly  fitting 


Fig.  9. 


the  cam  lever  into  the  slot  of  the  roller  head,  in  order  to  insure  proper  rela- 
tion of  roller  and  cam.  The  intentional  slanting  of  the  ribs  with  a  view  to  produ- 
cing continuous  rotation  of  the  valve  to  insure  better  tightness  has  shown  no 
advantage.  The  valve  spring  should  be  made  adjustable  in  order  to  overcome 

the  increased  friction  during  the 
first  period  of  operation.  Large 
bonnets  should  be  provided  with 
a  false  cover  or  be  otherwise  insu- 
lated from  the  live  steam  space, 
so  that  radiation  may  be  redu- 
ced, and  the  transfer  of  heat  to 
the  valve  gear  parts  minimized. 
Valve  roller,  pins  and  cam  lever 
should  be  made  of  steel,  hardened 
and  ground.  It  is  better  to  place 
the  valve  roller  on  the  roller  head 
than  on  the  cam  lever. 

The  radial  difference  between 
the  upper  and  lower  lands  of  the 
cam  corresponds  to  the  valve 
lift.  The  lifting  curve  should  rise 
tangentially  from  the  lower  land 
and  join  the  upper  land  by  means 

of  another  curve  or  fillet  which  runs  tangentially  into  it.  Sometimes  a  straight 
piece  is  inserted  between  the  two  curves,  with  the  object  of  reducing  the  high 
acceleration  at  the  point  where  the  curvature  changes. 


Calculation  of  the  Valve  Spring. 

The  purpose  of  the  valve  spring  is  to  keep  the  cam  and  roller  positively  in 
contact  during  the  time  the  valve  is  lifted  and  to  keep  the  valve  on  its  seat  while 
it  is  closed.  The  parts  to  be  accelerated  are  valve,  valve  stem,  roller  head,  roller, 
roller  pin,  and  valve  spring.  The  spring  also  has  to  overcome  the  steam  pressure 
upon  the  cross-section  of  the  valve  stem.  It  is  always  advisable  to  make  sure  of 
the  amount  and  direction  of  this  pressure,  since  there  are  cases  (exhaust  valves 
of  high  pressure  cylinders)  where  this  pressure  tends  to  relieve  the  valve  spring. 
The  maximum  accelerations  and  retardations  are  partly  produced  by  the  valve 
gear  and  partly  by  the  spring.  The  object  of  the  present  calculation  is  to  deter- 
mine the  maximum  acceleration  to  be  produced  by  the  valve  spring  and  the  neces- 
sary dimensions  of  the  latter.  The  usual  method  consists  in  plotting  the  valve 
lift  on  a  time  basis  and  differentiating  twice  to  obtain  a  curve  of  accelerations, 
from  which  the  maximum  acceleration  may  be  taken  and  the  spring  calculated. 
This  method  is  not  very  exact.  The  following  method,  based  on  the  same  prin- 
ciple, is  more  accurate. 

The  calculation  will  be  made  for  the  largest  cut-off  and  a  speed  of  100  r.  p.  m., 
the  angularity  of  the  eccentric  rod  being  neglected.  Then  the  angular  velocity 


89 


o>=  10.5,  and  the  time  for  1°  of  crank  angle  t  —  1/600  sec.    For  any  other  speed  n 


the  valve  velocity  v  must  be  multiplied  by  TT-T-  and  the  acceleration  by 

Time  or  crank  angle  is  taken  as  the  abscissa  for  all  calculations.   The  investi- 
gation will  extend  only  to  the  point  at  which  the  cam  lever  reverses,  since  the 


j-i 


Fig.  11. 


subsequent  closure  of  the  valve  is  an  exact  repetition  of  the  phases  during  lifting. 
The  necessary  dimensions  of  the  valve  gear  parts  are  assumed  to  be  known  or  are 
determined  according  to  Figs.  12,  13  and  14.  From  Figs.  13  and  14  the  travel 
of  the  cam  lever  (a  =  KK'}  can  be  found  corresponding  to  the  crank  angle  a. 


90 


\ 


Fig.  12. 


This  corresponds  to  the  part  MM'  of  the  valve  lift  curve  (equidistant  from  the 

cam  profile  through  the  center  of  the  cam  roller)  and  the  valve  lift  h.  The  cam 

lever  therefore  changes  a 
I  into  h.  The  line  / — K  in 

Figs.  12  and  13  corresponds 
to  the  z-axis  in  Fig.  14.  The 
distance  a  =  KK'  corresponds 
to  the  angular  displacement 
/?  of  the  cam  lever.  Imagine 
the  roller  to  move  along  the 
cam  (the  latter  supposed 
held  stationary)  from  M  to 
M',  corresponding  to  the 
angle  /5.  ML  is  the  center 
line  of  the  valve  stem. 
Drawing  M'  L'  tangentially 
to  an  arc  struck  from 
center  N  with  radius  N L, 
the  base  circle  MM'  will  be 
intersected  at  a  point  whose 
distance  from  M'  =h=  valve 

lift,   corresponding  to   a  and   p.    The  valve  lift  curve   M  M'Q  is  obtained  from 

the  cam  profile  by  increasing  or  decreasing  the  radius  of  the  latter  by  the  roller 

radius.     O^M  is  usually  from  3  to   10   mm,   and  O^M' 

=  roller  radius  -f-  3  to  10  mm.    The  cam  profile  as  well 

as  the  valve  lift  curve  consist  of  two  circular  arcs  with 

a  common  tangent  between  them.    As  long  as  the  roller 

remains   on   the    arc   of   the  cam   with  center  O^  it   is 

advisable  for  the  sake  of  accuracy  to  take  the  in- 
crements of  the  crank  angle  a  equal  to  2°,  while, 

later  on,  increments  of  5°  are  sufficient.     Fig.  15  shows 

the    valve    lifts    thus    determined    plotted    against    the 

crank  angle  a.  ^max  cor- 
responds to  the  dead  center 

position     of      the     eccentric. 

This  curve  is  also   important 

for   the  determination  of  the 

admission    line     (see    chapter 

on  losses    due  to   throttling). 

Instead  of  now  differentiating 

this  valve  lift  curve    to   find 

the  valve  velocities,  the  latter 

may  be  determined  in  a  more 

exact  manner  from  the  in- 
stantaneous angular  velocity 

and   the    corresponding    lever 


91 


arms.      The   velocity  of  the  eccentric  rod   H  K  in  Fig.  12  is  found  from  Fig.  14 
to  be 


£>!  =  V  -1  =  — 

r 


2  •  n  •  r  •  n 


bO 


=  co 


co  •  r-, 


The  angular  velocity   of   the  cam  lever  is  o^  =  -    —  (Fig.  13).  The  "roller 

r2 

contact  line"  M'02P  always  intersects  the  roller  center  and  either  of  the  centers 
O:  or  <92,  or  ig  at  right  angles  to  the  straight  middle  piece  of  the  valve  lift  curve. 
The  velocity  along  the  roller  contact  line  is  v'  =  o^  •  r3'.  For  the  velocity  along 
the  relative  valve  stem  axis  M'  L'  (valve  velocity  y),  7*3  has  to  be  considered  instead 
of  r3',  since  the  right-angled  triangles  NSP  and  M'  TU  are  similar,  and  v  :  v' 


=  r3  :  r3'.    Therefore  the  valve  velocity  v 


r3  =  co  •  —  •  r3.    The  distances 


r-,,  r9  and  r,  are  to  be  measured  in  meters.    The  roller  contact  line  M'0ZP  coin- 


cides with  MN  for  the  point  of  valve  opening 
and  therefore  rs  =  0.  The  same  occurs  when  the 
roller  reaches  the  upper  land  and  M'  falls  on  Q, 
in  which  case  also  r3  =  0.  The  valve  velocities 
thus  determined  are  plotted  in  Fig.  15  on  a 
crank  angle  basis,  giving  the  curve  y,  which 
indicates  a  rapid  increase  of  the  velocity  up 


IT 


Fig.  14. 


Fig.  15. 


to  the  point  7\,  corresponding  to  the  point  at  which  the  lifting  curve  and 
straight  portions  merge,  after  which  it  decreases  until  it  reaches  zero  for  the 
dead  center  position  of  the  eccentric,  or  for  the  point  at  which  the  upper  curve 
and  upper  land  run  together.  The  velocity  u  =.  0  corresponds  to  the  dead  center 
position  of  the  eccentric  after  which  the  whole  procedure  is  repeated  with  the 
signs  reversed.  In  case  the  roller  reaches  the  upper  land  or  runs  a  certain  distance 
on  it,  a  corresponding  part  of  the  velocity  curve  coincides  with  the  #-axis. 

The  maximum  valve  acceleration  p  =    —corresponds  to  the  steepest  tangent 

Wv 

7\  T2  of  the  velocity  curve.  Receding  from  point  Tz  a  distance  equal  to  6°  crank 
angle  or  1/100  sec,  an  ordinate  at  this  point  will  represent  the  change  of  velocity 
in  Vioo  sec-  In  this  case  the  latter  amounts  to  vx  =  0.205  m/sec.  The  accelera- 


92 

tion   is   consequently  p  =-       -  ==20.5  m/sec.    If  the  scale  of  velocity  is  chosen 

/ioo 

so  that  100  mm  =  1  m/sec,  then  the  ordinates  in  mm  give  directly  the  accelera- 
tions in  m/sec.  The  accelerating  force  during  the  period  of  increasing  velocity  is 
exerted  by  the  eccentric  through  the  cams,  and  for  decreasing  velocity  the  retarding 
force  must  be  produced  by  the  valve  spring.  The  opposite  holds  true  for  the 
closing  period  of  the  valve. 

The  force  to  be  exerted  by  the  valve  spring  depends  upon: 

1.  The  maximum  acceleration  pmax  to  be  produced  by  the  spring. 

2.  The  weight  G  of  the  valve  and  the  parts  connected  to  the  same. 

3.  The  friction  of  the  valve  stem,  valve  and  roller  head. 

4.  The  steam  pressure  upon  the  valve  stem  area. 

The  weight  G  is  generally  assumed  to  be  balanced  by  the  friction  and  there- 
fore neglected. 

g 

The  accelerating  force  P  =  jz-rrr  •  pmax.    An  additional  10  to  20%  are  neces- 

y.oj. 

sary  to  take  care  of  inaccuracies  in  the  construction  of  the  valve  gear  and  the 
increased  friction  during  the  first  period  of  operation.  The  steam  pressure  on 
the  valve  stem  area  may  either  relieve  or  oppose  the  spring  and  must  in  every 
case  be  considered.  Torsional  stress  of  the  spring  material  kd  =  about  3500  kg/sqcm. 


Pistons. 

Pistons  may  be  classified  as: 

1.  Self-supporting  pistons. 

2.  Floating  pistons. 

3.  A  form  intermediate  between  the  other  two. 

All  questions  concerning  the  design,  construction  and  operation  of  the  piston 
should  receive  thorough  consideration.  The  self-supporting  piston  is  undoubtedly 
the  most  difficult,  and  the  floating  piston  the  easiest  to  deal  with.  The  self- 
supporting  piston  together  with  its  cylinder  form  a  bearing.  The  first  condition 
for  satisfactory  operation  is  therefore  sufficient  difference  in  the  properties  of 
the  materials  of  which  the  two  parts  are  made.  For  instance,  a  steel  shaft  runs 
quite  satisfactorily  in  a  babbitt,  phosphor  bronze,  brass  or  cast  iron  bearing.  All 
these  combinations  present  sufficient  difference  in  the  properties  of  the  materials 
employed.  An  exception  exists  in  the  combination  of  hardened  steel  on  hardened 
steel  which  may  frequently  be  found  in  valve  gear  joints.  Babbit  on  babbitt,  bronze 
on  bronze,  or  mild  steel  on  mild  steel  never  work  together  satisfactorily.  Cast 
iron  pistons,  however,  can  be  made  to  work  well  in  cast  iron  cylinders,  as  is  shown 
by  the  una-flow  engines  built  by  Sulzer  Bros.  Cast  iron  is  a  collective  designation 
which  includes  materials  of  very  heterogeneous  composition.  Sulzer  Bros,  after 
extensive  experiments  have  found  the  proper  mixtures  for  the  cylinder  and  piston, 
which  possess  sufficient  difference  in  properties  to  work  together  safely  and  satis- 
factorily. If  a  designer  lacks  sufficient  faith  in  his  foundry  he  will  do  well  to  equip 
his  piston  with  a  babbitt  or  bronze  mounting  in  order  to  provide  materials  with 


93 

sufficient  difference  in  properties.  A  cast  steel  piston  should  always  be  equipped 
with  such  a  bearing  surface.  There  are  still,  however,  concerns  who  try  to  make 
piston  and  cylinder  from  the  same  mixture,  and  in  addition  to  this  cardinal  error 
commit  others  of  equal  consequence  which  result  in  certain  failure.  There  are  also 
materials  which  do  not  work  well  together  despite  sufficient  difference  in  their 
properties,  such  as  for  instance,  cast  steel  on  cast  iron.  Even  with  a  bronze  mounting 
the  difference  in  expansion  between  these  two  materials  must  be  taken  into  con- 
sideration. 

A  journal  and  bearing  must  ordinarily  have  sufficient  clearance  to  allow  room 
for  the  oil  film.  This  condition  applies  all  the  more  to  piston  and  cylinder  since 
the  dimensions  are  larger  and  the  temperature  difference  is  greater.  A  clearance 
of  3,5  to  4  thousandths  of  the  diameter  between  piston  and  cylinder  bore  have 
been  found  satisfactory.  Machining  the  piston  by  first  turning  it  to  the  exact 
cylinder  diameter  and  afterwards  turning  it  off  eccentrically  so  as  to  produce  a 
bearing  surface  for  about  90  to  120°  has  not  proved  satisfactory  for  superheated 
steam.  This  method  gives  enough  clearance  at  the  top  but  on  account  of  the  higher 
temperature  of  the  piston  and  its  greater  expansion,  the  weight  concentrates  at 
the  edges  of  the  bearing  surface  and  the  piston  is  likely  to  seize.  A  proper  way 
of  finishing  the  cylinder  is  to  bore  it  barrel-shaped,  or  machine  it  while  heating 
the  ends,  for  instance  by  passing  steam  through  the  jackets  and  cooling  the  center 
by  blowing  air  through  the  exhaust  belt.  The  heads  of  a  una-flow  piston  expand 
more  than  the  center  by  reason  of  their  higher  temperature,  and  should  therefore 
be  of  smaller  diameter  than  the  center.  They  should  be  out  of  contact  with  the 
cylinder  and  only  act  as  plungers.  The  part  of  the  piston  forming  the  bearing 
surface  should  be  rounded  off  liberally  at  its  ends  to  prevent  it  from  scraping  the 
oil  off  the  cylinder  wall.  Briefly,  the  endeavor  must  be  to  produce  a  piston  and 
cylinder  with  exact  cylindrical  surfaces  and  sufficient  clearance,  and  to  maintain 
this  condition  at  high  temperatures.  If  this  can  be  accomplished  for  the  long  una- 
flow  piston  with  its  large  bearing  surface  and  temperature  difference,  the  most 
difficult  part  of  the  problem  is  solved. 

The  case  of  the  floating  piston  which  is  carried  by  the  piston  rod  is  much 
simpler.  A  radial  clearance  of  2  to  3  mm,  according  to  the  size  of  the  cylinders, 
may  be  used,  so  that  no  consideration  of  bearing  action  is  necessary.  Only  the 
piston  rings  project  beyond  the  piston  surface  and  are  in  contact  with  the  cylinder 
wall.  The  tail  rod  can  either  have  a  stationary  bearing  behind  the  stuffing  box  or 
be  carried  on  a  crosshead.  The  great  length  of  the  piston  rod  between  the  bearings 
caused  by  the  long  piston,  makes  a  light  cast  steel  construction  and  a  rod  of  large 
diameter  a  necessity.  From  a  thermal  point  of  view  the  floating  piston  is  to  be 
preferred,  since  only  the  piston  rings  transmit  heat  from  the  hot  to  the  cold  por- 
tions of  the  cylinder,  while  in  the  case  of  the  self-supporting  piston  the  large  bearing 
surface  also  takes  part  in  this  action.  When  using  a  stationary  bearing  for  the 
tail  rod  behind  the  stuffing  box  there  will  be  a  rise  and  fall  of  the  piston  at  every 
stroke,  which  must  be  considered.  If  the  cast  steel  used  is  very  soft,  then  the 
cast  iron  rings  are  liable  to  seize. 

The  long  and  heavy  piston  together  with  the  long  span  of  the  rod  usually 
bring  about  a  condition  which  is  a  mean  between  the  floating  and  self-supporting 


94 

constructions.  Part  of  the  weight  is  then  carried  by  the  cylinder  wall  and  part 
by  the  piston  rod.  The  wear  of  the  cylinder  tends  to  alter  the  weight  distribution 
in  such  a  way  as  to  increase  the  part  carried  by  the  piston  rod  and  thus  relieves 
the  piston  and  cylinder  bearing  surfaces.  The  piston  rod  also  resists  possible 
forces  acting  on  the  outside  of  the  piston.  If  for  instance,  the  piston  rests  in  the 
cylinder  and  by  reason  of  leaky  rings  the  steam  obtains  access  to  the  clearance 
space  above  the  piston,  then  a  heavy  downward  load  will  result.  Floating  pistons 
offer  greater  safety  against  this  possibility,  this  safety  being  imparted  by  the 
piston  rod.  This  applies  especially  to  vertical  engines  where  the  piston,  if  not  guided 
by  a  tail  rod,  is  in  a  condition  of  unstable  equilibrium  and  is  liable  to  slap.  This 
may  even  be  occasioned  by  the  piston  overrunning  the  inlet  ports,  which  should 
be  fundamentally  avoided.  Such  overrunning  can  only  be  permitted  in  the  case 
of  floating  pistons,  although  even  then  the  possibility  of  vibration  should  be  reckoned 
with.  The  greater  number  of  mistakes  which  give  rise  to  lateral  forces  are  incurred 
in  the  arrangement  and  construction  of  the  piston  rings,  which  latter  should  pro- 
tect the  piston  surface  against  such  forces.  For  this  reason  they  should  be  placed 
as  far  as  possible  towards  the  ends  of  the  piston,  in  order  to  leave  as  little  area 
as  possible  for  the  formation  of  lateral  forces.  Even  if  this  is  done  there  still 
remains  some  possibility  of  an  unbalanced  load,  especially  with  the  large  surface 
of  a  una-flow  piston.  If,  for  instance,  the  rings  of  a  horizontal  engine  are  not 
secured  against  creeping,  their  center  of  gravity,  being  located  eccentrically  oppo- 
site the  joints,  will  move  to  the  lowest  possible  position  and  all  the  joints  will 
fall  in  line  at  the  top.  The  steam  leaking  through  them  will  then  undoubtedly  exert 
a  heavy  pressure  upon  the  large  surface,  thus  forcing  the  piston  downwards  and 
causing  rapid  wear.  This  cannot  happen  with  a  floating  piston  since  the  pressures 
will  equalize  in  the  annular  space  and  temporary  lateral  forces  will- be  resisted  by 
the  piston  rod.  The  ring  joints  of  floating  pistons  and  of  pistons  having  the  rings 
on  the  plunger  heads  should  therefore  be  equally  spaced  over  the  circumference; 
for  instance,  where  three  rings  are  used  at  each  end,  the  joints  should  be  at  120°. 
In  self-supporting  pistons  without  plunger  heads,  the  ring  joints  should  be  kept 
within  the  bearing  area;  thus  for  three  rings  one  joint  may  be  arranged  in  the 
center  and  one  each  at  30°  to  the  right  and  left.  The  bearing  surface  then  pro- 
tects the  ring  joints  against  the  steam.  With  such  an  arrangement  complete  tight- 
ness may  be  attained  if  the  workmanship  is  good  and  the  rings  are  sufficient  in 
number.  In  floating  pistons  the  action  of  the  rings  is  similar  to  a  labyrinth  packing 
which  always  passes  a  certain  amount  of  steam,  since  the  ring  joints  can  never  be 
made  absolutely  tight.  The  pressure  ratio, in  case  of  una-flow  engines  always  being 
above  the  critical  value,  the  weight  of  steam  flowing  past  the  ring  joints  may  be 
calculated  by  means  of  the  formula 


in  which  /  denotes  the  free  area  of  the  joint,  z  the  number  of  joints  in  series,  p^ 
and  vl  the  absolute  pressure  and  the  specific  volume  of  the  steam  inside  the 
cylinder.  In  Fig.  16  are  illustrated  five  different  types  of  ring  joint  fastenings 
which  can  only  partly  render  the  joints  tight,  but  perform  the  important  addi- 


95 


to 

£ 


96 


tional  function  of  securely  locking  the  rings  against  creeping.  In  a  una-flow  piston 
where  the  rings  are  usually  mounted  on  the  piston  heads  which  expand  conside- 
rably, the  locking  elements  must  in  no  case  project  to  the  cylinder  wall. 

The  slot  at  the  ring  joints  must  be  from  2  mm  to  5  mm  wide,  according  to 
the  diameter  of  the  rings,  in  order  to  allow  for  expansion.  Friction  will  cause 
the  rings  to  assume  a  higher  temperature,  especially  with  superheated  steam  and 
poor  lubrication.  If  the  clearance  provided  is  insufficient,  the  joints  will  close 
and  the  rings  expand  against  the  cylinder  wall,  so  that  increased  heating,  greater 
expansion  and  a  heavier  pressure  result,  which  may  lead  to  a  complete  destruction 
of  the  cylinder  surface  and  rings. 

Dimensions  of  Concentric  Cast  Iron  Piston  Rings  in  mm. 


Cylinder  Bore] 

Piston  Ring 

Length  of 
piece  cut  out 

Thickness 

Width 

300 

12 

12 

24  ' 

400 

15 

15 

35 

600 

20 

20 

GO 

800 

22 

22 

84 

1000 

25 

25 

108 

1200 

28 

28 

132 

1400 

30 

30 

155 

1600 

32 

32 

180 

1800 

34 

34 

205 

2000 

36 

36 

230 

The  different  phases  in  the  manufacture  of  ordinary  piston  rings  are  pre- 
sented in  Fig.  17.  First  is  shown  the  rough  casting  of  sufficient  width  to  hold  it 
in  the  lathe,  rough  turning  and  cutting  off  follow;  a  piece  is  then  cut  out  and  the 
ring  closed  up  for  finish  turning  to  the  correct  diameter.  It  is  not  possible  to 
obtain  a  uniformly  distributed  pressure  with  concentric  rings,  and  only  a  very 
rough  approximation  to  this  condition  can  be  reached  with  excentric  rings  whose 
thickness  increases  from  the  joint  to  the  opposite  side.  The  concentric  type, 
however,  is  preferable  in  order  to  avoid  a  one-sided  center  of  gravity  and  a  large 
clearance  in  the  groove  behind  the  ring. 

Attention  may  also  be  drawn  to  what  are  known  as  hammered  rings.  These 
are  made  of  Swedish  iron,  turned  to  correct  diameter  and  width  in  one  operation, 
are  then  split,  and  afterwards  given  the  required  tension  by  hammering.  Approxi- 
mately uniform  pressure  distribution,  greater  strength  and  reliability  in  operation 
are  obtainable  with  this  form.  Even  in  small  sizes  they  may  be  sprung  sufficiently 
to  be  slipped  into  the  piston. 

Mention  should  also  be  made  of  the  piston  rings  designed  by  Schmeck  (Fig.  18), 
which  are  made  in  sections  whose  joints  are  secured  by  spring-loaded  plugs  which 
prevent  them  from  creeping  and  force  them  against  the  cylinder  wall.  Rings  of 
this  type  have  the  advantage  that  they  are  practically  tight,  especially  at  the 
joints,  that  their  bearing  pressure  is  uniform,  they  can  be  easily  assembled  and 
disassembled,  and  adapt  themselves  to  warped  cylinders.  The  cast  iron  plugs 


97 

and  ring  sections  are  finished  in  such  a  way  as  to  fit  the  cylinder  bore.  This  type 
of  ring  is  much  used  by  the  Hannoversche  Maschinenfabrik  vorm.  Egestorff  and 
is  reported  to  have  given  complete  satisfaction. 


n. 


V,06S'£- 


i 

ro 


V5 


-19  60  f- 


Fig.  17. 

It  is  advisable  not  to  permit  the  outer  ring  to  overrun  the  cylinder  bore.  Such 
overrunning  exposes  part  of  the  ring  surface  to  the  steam  pressure,  thus  causing 
the  ring  to  collapse  and  destroying  its  function  of  tightness  for  at  least  a  certain 
distance  near  the  dead  center.  An  ordinary  cast  iron  ring,  especially  if  the  de- 


Fig.  18. 


flections  are  large,  will  not  withstand  the  stresses  produced  thereby  for  any  length 
of  time.  The  clearance  behind  the  rings  should  be  made  as  small  as  possible,  so 
as  to  reduce  the  deflection  and  leave  as  little  space  as  possible  for  the  accumula- 

Stump/,  The  una-flow  steam  engine.  7 


98 


tion  of  steam  which  would  press  the  ring  against  the  cylinder  wall  and  cause  con- 
siderable wear  both  on  the  ring  and  cylinder,  especially  in  the  middle  of  the  latter. 
With  superheat  and  dirt  in  the  steam  a  ring  may  under  these  conditions  lose  as 
much  as  5  to  10  mm  in  thickness  in  a  few  weeks,  and  the  cylinder  bore  several 
millimeters,  especially  in  the  center.  A  wide  ring  will  be  the  most  subject  to  this 

destructive  action.  Small  width,  high  grade 
material,  no  overrunning,  small  clearance 
behind  the  rings,  well  secured  joints,  and  a 
good  fit  in  the  grooves  are  therefore  advisable. 
If  the  cylinder  is  made  of  a  fairly  hard, 
close-grained  cast  iron,  then  no  ridges  will 
form,  thus  eliminating  any  reason  for  al- 
lowing the  rings  to  overrun. 

All  the  foregoing  is  supported  by  the 
experience  which  the  author  gained  with 
a  piston  packing  of  the  type  shown  in 
Fig.  19.  Both  rings  overran  the  cylinder 
bore.  The  rings,  in  collapsing,  had  to  push 

the  spring  along  their  sloping  surfaces,  with  the  effect  that  both  rings  and  spring 
went  to  pieces.  Fig.  20  shows  the  pieces  which  the  author  found  in  the  cavity  of  the 
piston.  The  spots  where  wear  occurred  on  the  springs  are  clearly  visible  in  Fig.  2Q. 

The  number  of 
rings  used  should 
vary  according  to  the 
pressure  range.  The 


Fig.  19. 


the  rings  at  the  ends 
of  a  una-flow  piston, 
with  both  sets  active 
during  the  first  part  of 
the  stroke,  fulfills  this, 
requirement  in  the 
best  possible  manner. 
A  una-flow  piston 
should  always  be  ma- 
de in  two  or  three 
parts,  held  together 
by  piston  rod  and  nut. 
The  castings  in  this 
case  are  simple,  light 

and  without  core  plugs,  which  will  be  especially  appreciated  in  the  case  of  cast  steel. 
.It  is  advisable  to  test  the  piston  with  water  pressure  if  the  foundry  cannot  be  relied 
upon.  A  cast  steel  piston  of  two-piece  construction  is  shown  in  Fig.  5,  page  77. 
Each  half  carries  a  bronze  shoe  fastened  to  it  with  copper  rivets,  covering  an  angle 
of  120°.  The  two  halves,  fitted  with  three  rings  each,  have  radial  clearance  over  their 
whole  circumference  and  are  made  as  light  as  possible  in  order  to  reduce  inertia. 


Fig.  20. 


99 

The  ring  joints  have  a  labyrinth  effect.  Fig.  40,  page  157,  shows  an  older 
piston  of  one-piece  design  having  grooves  fitted  with  Allan  metal  rings.  The 
latter  are  made  to  project  about  1  mm  above  the  surface  of  the  piston  when  new, 
and  during  the  first  period  of  operation  part  of  this  metal  is  transferred  to  and 
fills  out  the  pores  of  the  cylinder  surface.  Both  these  rings  should  be  placed  in 
the  middle  of  the  piston. 

Una-flow  pistons  for  locomotives  are  always  made  in  three  pieces,  i.  e.,  two 
heads  carrying  the  piston  rings,  having  clearance  all  around,  and  a  center  sup- 
porting piece.  It  was  formerly  customary  to  make  the  latter  of  hard  steel,  but 
Swedish  iron  is  now  used  with  better  results.  The  heads  which  are  made  of  cast 
or  forged  steel,  expand  considerably  under  the  action  of  superheated  steam.  This 
expansion  is  transmitted  to  the  center  piece  and  must  be  considered  in  designing 
the  latter.  In  order  to  reduce  the  weight  of  the  reciprocating  parts  to  a  minimum, 


Fig.  21. 

locomotive  pistons  are  always  made  without  tail  rods  and  have  given  satisfaction 
except  for  minor  troubles.  These  have  been  due  to  errors  in  the  composition  of 
the  material  and  to  disregard  of  the  effect  of  the  expansion  of  the  heads  upon 
the  center  supporting  piece.  The  experience  obtained  with  such  pistons,  as  well 
as  the  favorable  observations  of  Sulzer  Bros.,  indicate  that  the  problem  may  be 
solved  with  self-supporting  pistons,  provided  the  important  requirement  of  a 
reliable  lubrication  system  is  satisfied.  One  oil  feed  on  top,  and  one  each  in  or  below 
the  horizontal  plane  on  both  sides  of  the  cylinder,  every  feed  being  connected  to 
a  separate  force  feed  pump,  will  give  satisfactory  results  with  the  proper  kind  of 
oil.  The  pump  plungers  should  be  timed  to  deliver  oil  only  during  the  periods 
when  the  corresponding  orifices  are  covered  by  the  piston.  If  the  feeds  are  con- 
nected to  the  cylinder  ends,  carbonization  of  the  oil  is  to  be  feared;  and  if  the  oil 
is  fed  to  the  center  of  the  cylinder  at  the  time  of  exhaust,  it  is  likely  to  be  blown 
through  the  ports.  Both  these  conditions  lead  to  a  high  oil  consumption  which  is 
sometimes  complained  of  in  connection  with  una-flow  engines.  The  time  during 

7* 


100 


Q 

« 


which  the  piston  covers  the  oil 
feeds  is  longest  when  the  latter 
are  in  the  middle  of  the  cylinder, 
so  that  even  with  pumps  having 
a  continuous  feed  there  is  a  reaso- 
nable certainty  of  oil  being  car- 
ried between  the  rubbing  surfaces. 
In  order  to  avoid  losses  due  to 
the  exhaust,  it  is  permissible  to 
arrange  the  three  feeds  close  to 
one  side  of  the  exhaust  belt  in- 
stead of  in  the  center.  In  regard 
to  lubrication  also,  the  floating 
piston  offers  greater  safety,  since, 
if  correctly  constructed,  the  rub- 
bing surface  is  limited  to  the 
rings.  If,  however,  the  steam 
obtains  access  to  the  spaces  behind 
the  rings,  heavy  friction  and  large 
oil  consumption  may  result.  With 
self-supporting  pistons  the  sanje 
effect  may  be  caused  by  lateral 
forces. 

Everything  considered,  it 
may  be  stated  that  the  self- 
supporting  piston  requires  greater 
care  in  regard  to  design,  choice 
of  material,  lubrication  and  ope- 
ration, but  has  the  advantage  of 
not  requiring  a  tail  rod  with  its 
bearing  or  crosshead.  The  floating 
piston  has  greater  reliability  but 
is  more  complicated  and  increases 
the  floor  space  required.  Self- 
supporting  pistons  may  however 
be  made  to  operate  satisfactorily. 

Piston  Rod  Packings. 

Soft  packing  is  used  only  in 
small  and  cheap  engines,  while 
metallic  packing  is  the  rule  with 
steam  of  high  pressure  and 
superheat. 

The  Lentz  packing  shown  in  Fig.  21  consists  of  a  plurality  of  one-piece  cast 
iron  rings,  whose  number  varies  according  to  the  pressure  to  be  carried.    These 


0,  3 


.S    o> 
02    £ 


6X3 

s 


101 

rings  are  fitted  to  the  rod  and  work  between  the  ground  surfaces  of  a  corresponding 
number  of  housings,  thus  producing  a  labyrinth  effect.  The  individual  housings 
form  metal  to  metal  joints  and  are  pressed  against  the  bottom  of  the  packing 
space  by  means  of  the  outside  gland,  sufficient  clearance  being  provided  to  allow 
the  cast  iron  rings  to  move  laterally.  The  last  chamber  collects  the  water  of  con- 
densation so  that  it  may  be  drained  away.  With  pure  steam,  absence  of  dust, 
and  good  lubrication  (the  oil  being  preferably  forced  into  the  packing  under  pres- 
sure), satisfactory  results  should  be  permanently  obtainable.  The  one-piece  rings 
may  sometimes  be  found  unhandy  in  assembling. 

The  American  type  of  packing  shown  in  Fig.  22  is  better  in  this  respect, 
since  it  may  easily  be  assembled  and  disassembled  and  offers  greater  freedom 
in  the  design  of  adjacent  parts.  Each  ring  is  made  in  four  pieces,  the  two  oppo- 
site segments  with  their  babbitt  lining  being  pressed  against  the  piston  rod  by 


means  of  springs.  The  second  ring  of  similar  construction  is  set  at  90°  to  the  first 
one.  The  two  remaining  segments  of  each  ring  are  forced  by  springs  against  the 
other  segments.  Botli  rings  are  properly  fitted  to  the  housing  at  their  joints  and 
relatively  to  each  other.  The  whole  packing  system  is  held  by  means  of  axial 
springs  against  a  spherical  seat,  thereby  accommodating  itself  to  inclined  positions 
of  the  piston  rod,  while  any  lateral  movement  of  the  same  is  provided  for  by  the 
sliding  fit  of  the  rings  in  their  housings.  This  packing  is  occasionally  stated  to 
be  unsatisfactory  for  vacuum,  although  this  criticism  may  be  unjust.  The  Duplex 
packing  shown  in  Fig.  23  which  is  equipped  with  an  additional  set  of  conical  bab- 
bitt rings,  is  equally  satisfactory  for  both  vacuum  and  high  pressure.  Different 


102 


Fig.  24. 


f     C    Ct     f7 


Fig.  25. 


Fig.  26. 


103 


springs  are  supplied  for  various  pressures.    Good  workmanship  is  claimed  for  this- 
packing.    Pure  steam,   regular  and  ample  lubrication,  as  well  as  frequent  use  of 
the  drain  cocks,  especially  in  running  in,  are  essential  for  success. 

Packings  of   a  similar  type   of  Simplex  and  Duplex  construction  are  offered 
by  the  firm  of  Max  Dreyer  &  Co.,  of  Magdeburg. 

The  Proell  packing  (Fig.  24)  is  based 
on  a  similar  principle.  Each  cast  iron 
ring  is  cut  into  six  parts  which  "are  held 
together  by  means  of  a  coil  spring,  the 
joints  of  one  ring  being  staggered  in  relation 
to  those  of  the  adjacent  one.  A  pair  of 
such  rings  is  contained  in  each  housing, 
which  is  easily  removable  by  means  of  an 
internal  lip.  These  housings  come  together 
on  metal  to  metal  joints  and  thus  form 
chambers  in  which  the  rings  have  suffi- 
cient play  to  enable  them  to  move  late- 
rally. The  whole  packing  is  held  together 
and  against  the  bottom  of  the  packing 
space  by  the  outside  gland.  The  oil  is 
either  fed  onto  the  piston  rod  or  forced 
between  the  middle  rings.  This  packing, 
however,  does  not  accommodale  itself  to 
inclined  positions  of  the  piston  rod.  This 
packing  is  furnished  in  Simplex,  Duplex 
or  Triplex  forms,  according  to  the  number 
of  rings  employed. 

The  Kranz  packing  (Fig.  25)  made 
by  the  Elementenwerk  Kranz,  of  Ludwigs- 
hafen,  also  employs  cast  iron  rings  in 
pairs,  each  cut  into  three  parts,  sur- 
rounded by  sectional  primary  housings 
held  together  by  coil  springs.  These  pri- 
mary housings  are  radially  movable  between 
the  ground  surfaces  of  the  secondary  hou- 
sings, provision  being  also  made  for  axial 
expansion  of  the  former.  Oil  may  be  fed 
under  pressure  to  the  packing,  although 
it  is  claimed  that  a  drip  feed  to  the  rod 
its  satisfactory.  The  use  of  three  pairs  of 
rings  with  a  large  number  of  cavities 
results  in  a  thorough  labyrinth  effect.  The 
water  of  condensation  is  caught  by  a  further 
pair  of  rings  on  the  outside,  so  that  it 
may  be  drained  away. 


104 

The  packing  designed  by  Wilh.  Schmidt  (Figs.  26  and  27)  takes  care  of  in- 
clined positions  of  the  piston  rod  in  the  best  possible  manner.  The  packing  as 
a  whole  is  inserted  between  two  rings  having  spherical  surfaces  with  a  common 
center,  these  in  turn  being  held  between  two  flat  surfaces  so  that  both  lateral 
and  rotative  movements  are  rendered  possible.  A  deep  recess  insures  further 
flexibility  as  well  as  a  cooling  effect.  The  segmental  babbitt  rings  of  conical  sec- 
tion are  held  together  by  a  powerful  spring.  A  special  fitting  containing  a  felt 
ring  and  having  an  oil  connection  serves  to  lubricate  the  rod  as  well  as  to  keep 
out  dirt.  This  packing  has  been  very  successful  on  locomotives. 


105 


6.  Losses  due  to  Radiation  and  Convection. 

It  will  be  seen  from  a  comparison  of  the  una-flow  with  the  ordinary  com- 
pound counterflow.  engine  that  the  former  can  have  only  small  radiation  and 
convection  losses.  The  radiating  surface  of  the  counterflow  engine  with  its  two 
cylinders,  receiver  and  accessories  is  two  or  three  times  as  large  as  that  of  the 
una-flow  engine,  with  correspondingly  higher  losses.  The  loss  due  to  radiation 
of  the  una-flow  cylinder  is  very  small  compared  with  the  radiation  losses  of  the 
steam  pipe,  for  which  reason  the  latter  will  be  dealt  with  first.  Assuming  a  flow 
of  superheated  steam  at  a  very  small  velocity  through  a  pipe  having  a  length  of 
100  m,  then  the  steam  at  the  far  end  will  have  a  lower  temperature  and  a  cor- 
respondingly larger  specific  weight,  (yx  :  v2  =  Tl  :  Tz)  but  practically  the  same 
pressure.  The  highest  point  E  in  the  temperature-entropy  diagram  shown  in 
Fig.  2,  chapter  I,  3 a,  corresponds  to  the  state  of  the  steam  at  the  entrance  of 
the  pipe,  while  the  lower  point  Q  on  the  same  pressure  line  represents  the  state 
of  the  steam  at  the  far  end.  The  narrow  vertical  strip  EFWQE  below  the  part 
EQ  of  the  pressure  line,  extending  down  to  the  line  of  zero  temperature  ( —  273°), 
represents  the  total  amount  of  heat  lost;  but  as  the  heat  represented  by  the  area 
of  the  diagram  below  the  back  pressure  line  cannot  be  utilized,  the  actual  radia- 
tion loss  is  represented  by  the  strip  ER  VQE.  Insulating  and  lagging  the  pipe 
can  therefore  only  result  in  a  mere  reduction  of  this  loss.  A  further  radiation 
loss  occurs  in  the  cylinder  and  will  show  itself  in  a  very  slight  deviation  to  the 
left  of  the  vertical  adiabatic  line.  Insulating  the  cylinder  can  therefore  only  tend 
to  reduce  this  slight  loss.  Much  more  important  is  sufficient  insulation  around 
the  cylinder  heads,  which  really  form  part  of  the  steam  pipe.  The  live  steam  pipe 
in  high  grade  plants  is  always  covered  while  the  cylinder  is  provided  not  only  with 
insulation,  but  lagging  as  well.  The  latter  supplements  the  effect  of  the  insulating 
material  in  an  efficient  manner  and  may,  if  constructed  of  several  casings  one 
within  another,  with  a  bright  inner  surface,  entirely  take  its  place.-  All  flanges 
should  also  be  covered.  The  materials  used  for  insulating  steam  pipes  and  cylin- 
ders are  Kieselguhr,  asbestos,  magnesia,  cork,  or  glass  or  textile  waste.  The  more 
porous  the  material  is,  and  the  thicker  the  layer,  the  better  will  be  the  insulating 
effect.  Part  of  the  heat  is  lost  by  radiation  and  part  by  convection.  The  process 
is  so  complicated  that  mathematical  treatment  fails  completely  and  actual  tests 
have  to  be  relied  upon.  The  following  table  will  help  to  clear  up  matters. 


106 


Maximum  drop  in  temperature  per  1  m  length  of  pipe,  for  14  at.  gage; 

Outside 

steam  pipe  insulated,  without  lagging. 

diameter 

Thickness  of 

300  "C 

350»C 

400»C 

of  steam 

Insulation 

steam  temperature 

steam  temperature 

steam  temperature 

80 

60 

40 

80 

60 

40 

80 

60 

40 

mm 

mm 

m/sec 

m/sec 

m/sec 

m/sec 

m/sec 

m/sec 

m/sec 

m/sec 

m/sec 

60 

0.08 

0.11 

0.14 

0.10 

0.14 

0.18 

0.13 

0.19 

0.25 

108 

80 

0.06 

0.09 

0.12 

0.08 

0.12 

0.16 

0.11 

0.17 

0.22 

100 

0.05 

0.08 

0.11 

0.07 

0.11 

0.15 

0.10 

0.16 

0.20 

60 

0.06 

0.09 

0.11 

0.08 

0.12 

0.15 

0.10 

0.16 

0.20 

133 

80 

0.05 

0.08 

0.10 

0.07 

0.10 

0.13 

0.09 

0.13 

0.17 

100 

0.045 

0.07 

0.09 

0.06 

0.09 

0.11 

0.08 

0.12 

0.15 

60 

0.045 

0.075 

0.09 

0.06 

0.09 

0.12 

0.08 

0.12 

0.16 

159 

80 

0.04 

0.065 

0.08 

0.055 

0.08 

0.11 

0.07 

0.11 

0.14 

100 

0.035 

0.055 

0.07 

0.05 

0.075 

0.10 

0.065 

0.10 

0.12 

60 

0.04 

0.06 

0.08 

0.05 

0.075 

0.105 

0.07 

0.105 

0.13 

191 

80 

0.035 

0.05 

0.07 

0.045 

0.07 

0.09 

0.065 

0.95 

0.12 

100 

0.03 

0.045 

0.06 

0.04 

0.06 

0.08 

0.055 

0.85 

0.11 

60 

0.033 

0.05 

0.065 

0.045 

0.07 

0.09 

0.06 

0.09 

0.12 

216 

80 

0.03 

0.045 

0.06 

0.04 

0.06 

0.08 

0.055 

0.08 

0.11 

100 

0.025 

0.04 

0.05 

0.035 

0.055 

0.07 

0.05 

0.075 

0.10 

60 

0.03 

0.045 

0.06 

0.04 

0.06 

0.08 

0.055 

0.085 

0.11 

241 

80 

0.025 

0.04 

0.05 

0.035 

0.055 

0.07 

0.05 

0.075 

0.10 

100 

0.023 

0.035 

0.045 

0.03 

0.05 

0.06 

0.045 

0.07 

0.09 

60 

0.028 

0.042 

0.055 

0.038 

0.057 

0.075 

0.053 

0.08 

0.10 

267 

80 

0.023 

0.035 

0.046 

0.033 

0.05 

0.065 

0.045 

0.07 

0.09 

100 

0.02 

0.03 

0.04 

0.028 

0.042 

0.055 

0.04 

0.06 

0.08 

60 

0.025 

0.038 

0.05 

0.033 

0.05 

0.065 

0.045 

0.07 

0.09 

292 

80 

0.023 

0.034 

0.042 

0.03 

0.045 

0.060 

0.04 

0.06 

0.08 

100 

0.020 

0.028 

0.036 

0.025 

0.044 

0.050 

0.035 

0.053 

0.07 

60 

0.022 

0.033 

0.043 

0.03 

0.045 

0.06 

0.042 

0.06 

O.OS3 

318 

80 

0.020 

0.030 

0.028 

0.026 

0.039 

0.052 

0.038 

0.056 

0.075 

100 

0.018 

0.027 

0.034 

0.023 

0.034 

0.045 

0.034 

0.05 

0.068 

60 

0.02 

0.03 

0.04 

0.027 

0.041 

0.055 

0.037 

0.058 

0.073 

343 

80 

0.018 

0.027 

0.035 

0.024 

0.036 

0.048 

0.033 

0.05 

0.065 

100 

0.015 

0.023 

0.03 

0.022 

0.032 

0.041 

0.029 

0.043 

0.057 

60 

0.018 

0.028 

0.037 

0,025 

0.038 

0.055 

0.035 

0.053 

0.07 

368 

80 

0.016 

0.025 

0.032 

0.023 

0.034 

0.045 

0.031 

0.047 

0.061 

100 

0.014 

0.022 

0.028 

0.021 

0.031 

0.04 

0.027 

0.042 

0.054 

60 

0.016 

0.024 

0.032 

0.023 

0.034 

0,046 

0.031 

0.047 

0.062 

394 

80 

0.014 

0.021 

0.028 

0.02 

0.03 

0.04 

0.027 

0.041 

0.055 

100 

0.012 

0.018 

0.024 

0.017 

0.026 

0.034 

0.024 

0.036 

0.048 

60 

0.015 

0.023 

0.031 

0.022 

0.033 

0.044 

0.03 

0.045 

0.060 

420 

80 

0.0135 

0.020 

0.027 

0.010 

0.028 

0.037 

0.026 

0.039 

0.52 

100 

0.012 

0.018 

0.024 

0.016 

0.024 

0.032 

0.023 

0.035 

0.44 

107 

It  will  be  observed  from  this  table  that  the  heat  loss  for  average  thickness 
of  insulating  material  is  inversely  proportional  to  the  steam  velocity.  Higher  velo- 
cities of  course  result  in  smaller  diameter,  circumference  and  surface  of  the  steam 
pipe,  thus  also  reducing  the  heat  loss.  Higher  velocities  are  therefore  advisable 
up  to  the  point  where  the  throttling  losses  become  excessive  (See  chapter  I,  3 a, 
especially  Fig.  2).  Heavy  insulation  (80  to  100  mm  thick)  is  to  be  recommended. 
Assuming  an  outer  temperature  of  0.°,  then  the  heat  losses  increase  faster  than 
the  temperature  gradient,  according  to  the  law  of  Stephan  Boltzmann. 

If  the  steam  cylinder  is  regarded  as  a  pipe,  the  steam  may  be  considered  to 
flow  through  it  with  a  velocity  equal  to  the  mean  piston  speed.  Applying  the  above 
rule  of  the  inverse  variation  of  the  heat  loss  with  the  steam  velocity,  it  will  be 
found  that  in  view  of  the  lower  temperature  the  radiation  losses  of  the  cylinder 
must  be  extremely  small,  especially  as  the  oil  film  and  the  lagging  form  part  of 
the  insulation.  These  losses  in  fact  are  so  small  as  to  appear  negligible  in  com- 
parison with  the  other  losses. 


108 


7.  Losses  due  to  incomplete  Expansion. 

A  loss -of  diagram  area  within  the  limits  of  the  piston  stroke  is  caused  by  the 
fact  that  the  exhaust  begins  with  a  certain  exhaust  lead  or  distance  /„  before  dead 
center  (Fig.  1),  and  this  loss  increases  as  the  exhaust  lead  fv  and  terminal  expan- 
sion, pressure  pe  increase.  For  small  exhaust  lead  and  low  terminal  pressure  this 
loss  is  negligible.  Even  in  non-condensing  una-flow  engines  in  which  a  large  ex- 
haust lead  is  used  in  order  to  soften  the  exhaust  puffs,  the  loss  of  diagram  area 
within  the  limits  of  the  piston  stroke  is  insignificant  when  compared  with  the 
lost  work  represented  by  the  toe  of  the  diagram,  shown  shaded  at  D.  The  pro- 
blem of  finding  a  means  of  utilizing  this  work  without  increasing  the  cylinder 
dimensions  is  well  worth  while.  The  solution  to  be  described  later  has  the  effect 
of  reducing  the  pressure  pu  at  which  compression  begins,  with  a  consequent  lower 


Fig.  1. 

terminal  pressure.  A  smaller  clearance  volume  may  therefore  be  used,  thus  dimi- 
nishing the  volume  loss.  Without  going  into  calculations,  it  will  be  seen  that  the 
gain  F  at  the  compression  line  will  be  approximately  proportional  to  the  shaded 
area  D,  or  in  other  words,  the  higher  the  terminal  expansion  pressure  is,  or  the 
longer  the  cut-off,  the  lower  will  be  the  terminal  compression  pressure.  This  implies 
an  increasing  pressure  difference  during  compression  for  an  increasing  pressure 
difference  during  expansion,  a  combination  which  has  been  proved  desirable  in 
the  chapter  on  volume  loss.  This  rule  would  be  fulfilled  in  its  entirety  if  the  pres- 
sure changes  on  both  sides,  i.  e.  expansion  and  compression,  were  equal.  Further- 
more, the  use  of  a  longer  exhaust  lead  /„  would  then  become  permissible,  since  the 
lost  area  within  the  limits  of  the  piston  stroke  now  forms  part  of  the  toe  of  the 
diagram,  and  co-operates  in  lowering  the  back  pressure  at  the  time  compression 
begins.  There  can  therefore  be  no  objection  to  making  fv  large,  since  by  increasing 
the  duration  of  the  exhaust,  the  compression  is  shortened  and  the  exhaust  puffs 
are  softened.  If  this  is  done,  the  number  of  the  exhaust  ports  will  be  so  far  reduced 
that  the  exhaust  belt  may  eventually  be  dispensed  with  and  only  one  port  remains 
which  connects  directly  to  the  exhaust  pipe.  Piston  and  cylinder  also  become 


109 


considerably  shorter.    The  relation  between  length  of  cylinder  L,  length  of  piston 
4,  exhaust  lead  /„,  stroke  /,  and  exhaust  port  diameter  d,  is  as  follows: 


Jt  =  d  +  I  —  If, 
L  =  2 


For  instance,  for  a  stroke  of  660  mm,  an  exhaust  lead  of  25%  and  exhaust 
ports  of  120  mm  diameter,  the  piston  length  will  be  450  mm,  and  the  length  of 
the  cylinder  1100  mm  as  compared  with  a  piston  length  of  594  mm  and  a  cylinder 
length  of  1254  mm  for  10%  exhaust  lead.  The  distance  fv  is  limited  by  the  maxi- 
mum cut-off,  since  direct  exhaust  of  live  steam  must  be  avoided.  For  locomotives 
the  value  of  fv  must  be  limited  to  25%,  while  for  locomobiles  it  may  be  taken  as 
large  as  30  to  35%. 

The  utilization  of  the  energy  represented  in  the  toe  of  the  diagram  is  based 
upon  its  complete  conversion  into  kinetic  energy  by  means  of  conical  nozzles 
such  as  are  commonly  used  in  steam  turbines.  Each  exhaust  puff  would  therefore 
act  as  a  kind  of  wad  or  plug  moving  with  a  high  velocity  through  the  exhaust 


ru* 


Fig.  2  (Diise  =  Nozzle). 

pipe  and  finally  creating  behind  itself  a  partial  vacuum  whose  absolute  pressure 
is  pu.  The  exhaust  pipe  must  therefore  be  long  enough  so  that  there  will  always 
be  at  least  one  such  plug  moving  within  it,  thus  preventing  atmospheric  pressure 
from  reaching  the  nozzle  and  destroying  the  vacuum.  The  end  of  the  exhaust 
pipe  must  form  a  diffusor  to  change  the  kinetic  energy  into  pressure  energy  at 
atmospheric  pressure. 

Fig.  2  shows  such  an  exhaust  pipe  diagrammatically.  The  work  corresponding 
to  the  shaded  areas  D  and  E  (Fig.  1)  is  determined  for  various  pressures  pu.  The 
weights  of  steam  Ge  and  Gu,  corresponding  to  the  pressure  pe  and  pu  can  be  found 
from  the  diagram  and  the  dimensions  of  the  engine. 

Using  the  equation  of  work: 

Gf—G,,    w2 


The  velocities  w  are  calculated  for  various  values  of  the  pressure  pu  and  plotted 
against  the  latter,  as  shown  in  Fig.  3.  Each  value  of  pu  has  associated  with  it  a 
certain  velocity  wd  which  must  exist  at  the  point  of  entrance  into  the  diffusor  in 
order  that  atmospheric  pressure  may  be  overcome.  This  velocity  therefore  cor- 
responds to  a  pressure  difference  (1  —  pu)  and  may  be  easily  obtained  from  the 
Mollier  chart  and  also  plotted  in  Fig.  3.  Of  course  the  steam  when  leaving  the 
diffusor  must  in  practice  still  have  a  certain  velocity,  and  the  pressure  difference 
should  therefore  be  reckoned  not  from  the  atmosphere,  but  from  a  slightly  higher 
pressure  corresponding  to  this  velocity.  This  shifts  the  diffusor  velocity  curve 


110 


into  a  somewhat  higher  position,  and  the  intersection  S  of  this  curve  with  the 
nozzle  velocity  curve,  which  determines  the  obtainable  pressure  pu,  is  moved 

slightly  towards  the  right  corre- 
sponding to  a  higher  value  of  pu. 
Friction  losses  in  the  long  exhaust 
pipe  cause  a  further  loss  of  velo- 
city «>!,  between  the  nozzle  and 

— -^>.  ~  the  diffusor,    and   this   again  re- 

x^^; — ^i   fit"  suits   in    a  shift   of  the  point  of 

r\^4Jlo7~""  intersection    to    the    right,    cor- 

responding to  a  still  higher  pres- 
sure pu.  The  ejector  effect  of 
the  exhaust  puffs  will  eventually 
become  less  and  less  for  higher 
steam  velocities.  This  will  be  the 
Fig.  3.  (Duse  =  Nozzle).  case  especially  in  single  cylinder 

engines,    because    the    length    of 

exhaust  pipe  required  is  very  great.  For  instance  in  an  engine  running  at  180  r.  p.  m. 
which  corresponds  to  6  exhaust  puffs  per  second,  and  for  a  steam  velocity  o- 
540  m/sec,  the  length  of  the  exhaust  pipe  must  be  90  m,  or  better  100  m.  Calf 
culating  the  loss  of  pressure  required  to  overcome  the  resistances  by  means, of 
Eberle's  equation:  , 


t.o 


in  which  A  p  represents  the  loss  of  pressure  in  kg/sqm,  I  is  the  length  of  the  ex- 
haust pipe  =  100  m,  d  its  diameter  =  0,1  m,  w  the  steam  velocity  =  540  m/sec, 
y  the  specific  weight  =  0.58  kg/cbm  and  $  a  constant  =  10.5  x  10~4,  then  the 
loss  of  pressure  will  be  found  to  be  11.1  at.  For  d  =  0.05  m  the  loss  would  be 
35.5  at.  However,  Eberle's  tests  from  which  the  above  formula  was  obtained, 
covered  only  velocities  up  to  150  m/sec,  as  did  similar  tests  by  Fritsche,  Ombeck, 
Lorenz  and  Ritschel,  so  that  the  results  are  not  directly  applicable  to  velocities 
higher  than  the  critical  value.  Such  higher  velocities  will  result  in  still  greater 
losses.  In  reality  the  above  results  will  be  smaller  since  the  assumed  steam 
velocity  will  not  be  constant  but  decreasing. 

It  would  seem  from  the  foregoing  that  the  direct  exhaust  ejector  principle 
would  not  have  great  prospects  if  based  on  complete  conversion  of  pressure  energy 
into  velocity.  However,  as  reported  by  Giildner2),  partial  vacua  have  been  observed 
in  long  exhaust  pipes  of  gas  engines,  although  no  special  provision  had  been  made 
to  cause  and  sustain  them  and  a  great  part  of  the  energy  was  lost  in  valves,  sharp 
edges  and  elbows.  There  must  therefore  still  remain  a  possibility  of  solving  this 
problem  in  another  way.  It  is,  however,  not  easy  of  solution,  since  it  involves 
the  calculation  of  the  friction  of  accelerating  and  expanding  steam,  which  is  very 
difficult  to  express  in  a  mathematical  form.  Actual  tests  must  therefore  be  relied 
upon. 


x)  Stodola,  ,,Die  Dampfturbinen".     3d.  Edition,  p.  55. 

2)  Giildner,  ,,Entwerfen  von  Verbrennungskraftmaschinen.     3d.  Edition,  p.  40. 


Ill 


Although  this  problem  may  seem  difficult  in  connection  with  a  single  cylinder, 
it  is  very  simple  for  multi-cylinder  engines,  of  which  the  locomotive  is  the  chief 
representative.  Even  in  an  engine  having  two  cranks  at  right  angles,  an  exhaust 
lead  of  25%  will  produce  sufficient  overlap  of  the  exhaust  periods  so  that  the 
exhaust  of  one  cylinder  begins  before  the  other  has  ceased  (Fig.  5).  If  now  the 
exhaust  pipes  are  joined  at  an  acute  angle,  a  jet  ejector  action  is  obtained.  This 
effect  is  well  known  and  widely  used  in  locomotive  practice,  where  the  combina- 
tion of  blast  pipe  and  stack  also  form  an  ejector,  thus  producing  a  partial  vacuum 
in  the  smoke  box  which  serves  to  draw  off  the  flue  gases.  The  theory  of  the  blast 
pipe  was  first  developed  by  Zeuner  in  his  classical  treatise  on  the  subject1).  An 
equation  in  a  somewhat  simpler  form  based  on 
this  theory,  is  found  in  v.  Ihering's  book,  ,,Die 
Geblase".  The  development  of  the  formula 
for  the  ratio  G2  :  G  of  the  quantities  of  the 
ejected  to  the  ejecting  steam  is  very  lengthy 
and  will  not  be  repeated  here.  The  formula  is 
based  upon  the  principle  of  the  continuity  of 
flow  and  includes  a  number  of  assumptions 
and  simplifications,  the  most  important  of 
which  are  that  friction  losses  are  not  con- 
sidered, but  the  unfavorable  assumption  is 
made  instead  that  the  velocity  wz  is  entirely 
lost  and  w3  =  0.  (See  Fig.  4.)  The  following 
formula  then  results: 


m 


m 


m  ^  n2 
2 


I 


u. 


in  which 


72 


The    efficiency    of 
upon    the    losses    due 
as    the    impact    losses 
the  two   streams  mix. 
regarded    as    absolutely 


ejector 
friction 


depends 
as    well 


an 
to 

at    the    point    where 

The  impact  is  to  be 

inelastic.     Assuming 


*ft 


Fig.  4. 

F  sectional  areas      w  velocities 
p  pressures  •/  specific  weight 

These  symbols  used  without  index  refer  to 

the  junction  of  the  nozzles  and  the 

steam  at  that  point; 
Index  2  refers  to  the  variable  port  opening 

at  the  cylinder; 
Index  1,  to  the  throat  of  the  blast  pipe  and 

the  mixture; 

Index  0,  to  the  final  section  of  the  blast  pipe; 
w3  =  the  velocity  of  the  ejected  steam  at  the 

junction. 


W 


=  0  then  the  efficiency  r]s  =  i 


Since  G2  is  small  compared  with 


G+G2 

the  ejecting  stream  loses  little  of  its  energy,  the  impact  loss  is  small  and 
Tis  =  0.75  to  0.80.  The  efficiency  of  the  blast  pipe  is  considerably  lower,  being 
only  about  0.28,  because  the  weight  of  the  ejected  air  is  about  2.6  times  the  weight 
of  the  ejecting  steam.  In  addition  to  this  there  is  a  further  loss  equal  to  the  energy 
contained  in  the  steam  at  the  final  section  of  the  stack.  A  similar  loss  will  occur 
at  the  end  of  the  blast  pipe  in  regard  to  the  exhaust  ejector  action,  since  the 
energy  contained  in  the  steam  at  the  final  section  P\  cannot  be  utilized  for  steam 
ejecting.  This  finds  its  expression  inequation  1,  where  A  increases  with  decreasing 


*)  Zeuner,  ;,Das  Lokomotiven-Blasrohr".     1863.     Zurich.     Meyer-Zeller. 


112 


F0.  The  area  F0  should  therefore  be  made  as  large  as  possible,  but  is  limited  by 
considerations  in  regard  to  the  blast  action  on  the  flue  gases.  Strahl1)  has  deve- 
loped a  formula  for  the  blast  pipe  area,  which  is  based  on  Zeuner's  treatise  and  is 

a-  R 
as  follows :  F  =  -y=,  where  F  is  the  blast  pipe  area,  F^  the  smallest  stack  area, 

y  xA 

R  the  grate  area,  A  the  coefficient  of  divergence  of  the  stack,  x  the  coefficient  of 
flue  gas  friction  from  ash  pan  to  stack;  a  =  f  (m);  m  =  A  •  F:  :  F  -  a  is  nearly  con- 
stant, and  S  0.03  for  m  =  13  to  19.  The  weight  ratio  of  the  ejected  air  L  to  the 
ejecting  steam  D  is 

L 

~D  = 


A  large  blast  pipe  area  produces  a  lower  pressure  in  the  cylinder  but  insuf- 
ficient vacuum  in  the  smoke  box  and  therefore  unsatisfactory  steaming  of  the 
locomotive.  It  is  self-evident  that  with  a  given  amount  of  work  available  in  the 

toe  of  the  diagram  only  a  certain 
total  of  ejector  action  for  cylinder 
and  boiler  together  can  be  produced, 
the  distribution  of  which  depends 
essentially  upon  the  blast  pipe  #rea. 
In  addition  to  the  blast  pipe  area, 
the  dimensions  of  the  stack  are  also 
important.  Too  small  a  stack  area 
leads  to  a  large  velocity  loss  at  the 
outlet,  and  too  large  an  area  causes 
a  considerable  impact  in  ejecting 
the  flue  gases.  It  follows  from  this 
that  there  must  be  a  best  stack 
area  for  which,  in  consequence  of 
the  losses  being  a  minimum,  the 
blast  pipe  area  is  a  maximum.  Ex- 
pressed mathematically,  a  =  f  (m) 
is  a  maximum  for  m  =  15.5  approxi- 
mately, as  is  demonstrated  in  the 

above-mentioned  paper  by  Strahl.  These  best  stack  dimensions  must  be 
strictly  adhered  to  in  a  locomotive  in  which  the  ejector  effect  of  the  exhaust 
is  utilized,  and  this  leads  to  considerable  difficulties  in  large  locomotives  of 
the  present  day.  The  length  of  the  stack  is  limited  to  such  an  extent  by  the 
loading  gage  and  the  high  location  of  the  boiler,  that  the  expansion  of  the  jet 
leaving  the  blast  pipe  may  not  completely  fill  out  the  whole  area  of  the  stack, 
so  that  air  may  enter  from  above  through  the  remaining  area  and  thus  partly 
destroy  the  vacuum.  Actual  tests,  however,  have  shown  that  such  loss  of 
vacuum  can  easily  be  avoided  even  with  a  large  stack  area. 

a)  Strahl,   ,,Untersuchung  und  Berechnung  der  Blasrohre   und  Schornsteine  an  Loko- 
motiven,  1912'',  Wiesbaden,  C.  W.  Kreidel. 


Fig.  5. 


113 


The  further  calculations  are  based  on  a  0—10 — 0  freight  locomotive  of  the 
German  State  Railways  as  an  example,  which  is  described  on  page  242.  This 
engine  is  a  two  cylinder  superheater  locomotive  with  a  cylinder  bore  of  630  mm, 
a  stroke  of  660  mm,  a  driving  wheel  diameter  of  1400  mm  and  a  steam  pressure  of 
12  at.  gage.  Assumed  is  an  evaporation  of  7000  kg/hour,  a  loss  of  pressure  of  1  at. 
from  boiler  to  valve,  adiabatic  expansion  and  compression  at  an  entropy  of  1.7, 
and  three  different  speeds  of  20,  40  and  60  km/hr,  which  are  referred  to  below  as 
cases  I,  II  and  III.  The  exhaust  port  of  the  cylinder,  having  a'diameter  of  120  mm, 
is  placed  7  mm  off  center  to  compensate  for  the  angularity  of  the  connecting  rod 
of  2600  mm  length.  The  diagram  shown  in  Fig.  5  is  based  on  these  figures  and 
the  cross-shaded  areas  represent  the  periods  of  overlapping  exhaust.  The  ejector 

action  is  effective  only  during  part  of  the  exhaust 
period  and  therefore  only  the  part  A  of  the  shaded 
area  in  Fig.  6  can  be  utilized  for  ejector  action,  the 
part  B  being  lost.  An  increase  in  the  exhaust  lead 
would  of  course  have  the  effect  of  increasing  A  by 
a  part  or  the  whole  of  B;  but  considerations  of 
evenness  and  strength  of  the  draft  forbid  this.  Of 
the  total  work  represented  by  the  area  A,  a  part 


Fig.  6. 

is  lost  in  producing  the  draft  in  the  smoke  box  by  the  rush  of  the  steam  at 
high  velocity  from  the  nozzle  (blast  nozzle  loss)  and  a  part  is  lost  by  throttling 
and  friction  in  the  pipes  (pipe  loss).  Finally,  after  subtracting  the  stack  loss, 
there  remains  the  effective  gain  of  work  C  at  the  compression  line,  corresponding 
to  an  absolute  pressure  pu. 

The  investigation  begins  with  the  determination  of  the  free  exhaust  port 
areas  for  various  piston  positions  within  the  limits  of  the  exhaust  lead  /„.  The 
secondary  nozzle  delivering  steam  to  the  feed  water  heater  must  also  be  included 
in  this  consideration,  since  the  steam  passing  through  it  acts  in  the  same  way 
as  in  the  main  exhaust  pipe.  Fig.  7  illustrates  the  profiles  of  the  main  and  secon- 
dary nozzles,  as  well  as  the  piston  positions  at  crank  angle  intervals  of  5°.  In 
order  to  find  the  smallest  areas  of  opening  of  the  nozzles,  the  outlines  of  which 
are  shown  shaded  in  Fig.  7  for  55°  before  dead  center,  their  plan  projections  are 
first  laid  out  and  their  areas  multiplied  by  Vcos  «•  In  calculating  the  weight  of 
steam  flowing  through  the  nozzles,  the  velocity  loss  must  be  taken  into  considera- 
tion by  using  a  velocity  coefficient  y  =  0.9.  On  account  of  the  very  unfavorable 
nozzle  profiles  when  first  uncovered,  with  consequent  turbulence,  it  was  consi- 
dered necessary  to  use  a  smaller  value  of  <p  at  first,  and  therefore,'  until  the  nozzle 
was  fully  open,  (p  was  put  =  0.7  to  0.9.  As  is  shown  in  Fig.  8,  the  full  opening 

Stumpf,  The  una-flow  steam  engine.  8    m 


114 


is  very  soon  reached,   since  the  piston  overruns  the  nozzle  by  a  considerable 

amount. 

For  the  several  piston  positions  at  intervals  of  5°,  the  weights  of  steam  passing 

through  were  calculated  by  means  of  the  equation  G  =  <p-  Fm-  w  •  y  - 1,  where  Fm 

denotes  the  smallest  mean  area  of 
opening  of  main  and  secondary  nozzles 
combined.  The  other  quantities  were 
taken  at  their  initial  values,  since  they 
can  vary  but  little  in  the  small  time 
intervals  concerned.  In  this  equation 
w  —  3.35  \  p  v  denotes  the  exhaust  ve- 
locity, where  p  is  the  steam  pressure, 

y  =  _  is   the    specific  weight   of    the 
v 

steam,  and  t  the  time.  When  G  kg 
of  steam  have  been  exhausted,  the 
piston  has  moved  on  and  the  cy- 
linder volume  V  has  reached  a  dif- 
ferent value,  which  is  tabulated  in 
Fig.  7.  By  dividing  the  new  cylin- 
der volume  by  the  weight  of  'the 
remaining  steam,  the  specific  volume 
in  the  new  piston  position  is  found, 
thus  determining  the  new  pressure  at 
this  point. 

The  velocities  vary  considerably 
below  the  critical  pressure  of  p  =  1,73 
at.  abs. ;  the  piston  however  has  then  fully  uncovered  the  nozzles,  and  Fm  is 
therefore  constant. 

For  this  period,  until  atmospheric  pressure  is  reached,  the  time  elements  or 
crank  angles  were  determined  by  the  methods  developed  in  Chapter  I,  3  a,  for 

finding  the  losses  due  to 
throttling.  The  results  in- 
dicated that  for  small 
velocities  the  area  was 
more  than  ample,  and 
that  for  higher  velocities 
the  exhaust  would  end 
at  atmospheric  pressure 
Fig.  8.  without  the  assistance 

of     the     ejector     effect. 

It  must,  however,  be  considered  that  all  this  refers  to  a  freight  locomotive  which 
ordinarily  runs  at  speeds  of  20  to  40  km/hour  and  that  the  back  pressures  re- 
maining in  the  cylinder  in  case  III  are  only  slightly  above  atmosphere.  A  con- 
siderable time  is  required  to  exhaust  the  last  tenth  of  an  atmosphere  on  account 
of  the  low  pressure  difference  and  the  small  exhaust  velocity.  The  expansion 


Fig.  7. 


-50   -SS  -So 


s  >«o0 


115 

and  exhaust  lines  I,  II  and  III  shown  in  Fig.  9  were  obtained  point  by  point  in 
this  manner,  corresponding  to  speeds  of  20,  40  and  60  km/hour.  The  part  of  the 
exhaust  line  where  the  ejector  action  comes  into  effect  was  also  determined  point 
by  point,  the  weights  of  steam  ejected  G2  being  calculated  by  means  of  equa- 
tion (1). 

As  long  as  the  pressure  in  the  cylinder  is  above  the  critical  value  pk  =  1,73  at. 
abs.,  the  pressure  energy  can  only  be  completely  converted  into  kinetic  energy 
by  the  use  of  conical  nozzles;  otherwise  the  jet  will  still  possess  some  pressure, 
and  the  suction  effect  will  be  diminished  on  account  of  the  lower  velocity  conse- 
quent on  the  decrease  in  specific  volume.  In  equation  (1)  this  is  taken  care  of 
by  an  increase  in  y.  The  theoretical  as  well  as  the  actual  factors  of  divergence 
i^  and  if}'  were  therefore  calculated  according  to  the  rules  of  steam  turbine  design, 


I 

I 

E 

J3mi 

6,8* 

S.95 

2.»i 

Fig.  9. 

and  are  also  tabulated  in  Fig.  7.  According  to  these  figures  the  actual  divergence 
is  insufficient  in  case  I  for  crank  angles  of  — 50°  to  — 30°.  A  correction  was  there- 
fore made  in  the  calculated  values  of  G2.  In  case  II  the  actual  divergence  is  al- 
most, and  in  Case  III  exactly  correct,  so  that  G2  need  not  be  corrected.  The 
compression  lines  in  Fig.  9  were  laid  out  in  accordance  with  these  results,  and 
it  is  found  that  the  initial  compression  pressures  are  respectively  0.7,  0.97  and 
1.0  at.  abs.  for  cases  I,  II  and  III. 

The  very  small  pressure  reduction  in  cases  II  and  III  makes  it  desirable  to 
analyze  the  losses  during  the  exhaust  period.  The  blast  pipe  loss  can  easily  be 
calculated  since  we  know  the  weights  of  steam  exhausted  during  the  time  the 
piston  travels  the  distances  fd  and  /f,  as  well  as  the  blast  pipe  area  and  the  specific 
volume  corresponding  to  atmospheric  pressure.  With  these  data  the  mean  steam 
velocity  at  the  outlet  of  the  blast  pipe  and  the  corresponding  energy  may  be 
calculated.  To  be  exact,  instead  of  taking  the  mean  velocity  w,  it  would  be  neces- 

8* 


116 


sary  to  find  the  value  of  the  integral  §w*dd  but  for  a  rough  approximation  this 

o 

is  not  necessary.  The  amount  of  work  represented  by  the  areas  A,  B  and  C  may 
be  found  with  a  planimeter,  and  the  efficiency  of  the  ejector  is  determined  by 
the  ratio  of  G2  :  G,  so  that  the  remainder  represents  the  pipe  loss.  The  major 
part  of  the  latter  is  caused  by  the  throttling  of  the  steam  when  entering  the  nozzle. 
A  high  vacuum  is  formed  in  the  exhaust  pipe  just  before  and  after  beginning  of 
compression  but  it  only  partly  reaches  the  cylinder.  This  is  taken  into  consideration 
in  equation  (1)  by  taking  a  lower  value  of  n.  The  following  table  gives  the 
relative  amounts  of  the  different  losses. 


Case 

I 

II 

III 

Area  A: 
Blast  nozzle  loss     

40 

25 

14 

Pipe  Loss  

14 

30 

56 

Impact  loss  during  ejection    . 
Gain  of  compression  
Area  B: 
Blast  nozzle  loss     

8 

27 

1 

4 

15 

9 

0 
0 

22 

Pipe  loss    

10 

17 

8 

Total  .    . 

100 

'  100 

100 

It  will  be  seen  from  this  table  that  the  blast  nozzle  loss  (part  A  and  B  taken 
together)  is  approximately  constant  and  amounts  to  from  41  to  36%.  The  pipe 
loss  increases  from  24%  at  low  speeds  to  44%  at  high  speeds,  and  the  work  re- 
presented by  area  B  from  11%  to  30%  respectively.  It  therefore  follows  that 
at  high  speeds  the  bad  effect  of  area  B  must  be  eliminated,  and  this  can  be  easily 
accomplished  with  a  three  cylinder  locomotive.  In  such  an  engine,  having  cranks 
at  120°  and  an  exhaust  lead  of  25%,  there  will  always  be  two  cylinders  exhausting 
at  the  same  time.  The  ejector  effect  begins  at  the  dead  center  and  proceeds  with 
greater  nozzle  areas  in  the  cylinder  wall,  with  the  result  that  the  throttling 
loss  and  pipe  loss  are  also  reduced.  The  loss  of  work  may  therefore  be  divided 
in  all  three  cases  as  follows: 

Blast  nozzle  loss 40% 

Pipe  loss   . 24% 

Impact  loss  ....*....     8% 

Useful  work.    .    .    ...    .    .   28% 

Assuming  this  division  of  losses,  and  taking  the  mean  effective  pressures  from 
the  indicator  cards,  the  following  figures  were  obtained  for  two  and  three  cylinder 
locomotives. 

The  three  cylinder  locomotive  also  shows  only  a  surprisingly  slight  gain  due 
to  the  ejector  effect  at  high  speeds  or  early  cut-offs.  This  is  explained  by  the 
fact  that  the  utilization  of  the  toe  of  the  diagram  is  equivalent  to  an  enlargement 
of  the  cylinder.  If  the  cut-off  is  early,  then  a  further  increase  in  expansion  will 
not  produce  much  gain.  The  steam  consumption  figures  are  already  so  low  that 


117 


I 

II 

III 

Two  Cylinder  Locomotive. 
Mean    effective     pressure,     with    ejector    effect, 
kg/sqcm 
Mean  effective  pressure   without  ejector 
effect     kg/5qcm 

6.84 
6.1 

3.93 
3.83 

2.83 
2.83 

Gain  due  to  ejector  effect  °/o 

12.2 

2.6 

Indicated    HP 

940 

1080 

1170 

Steam  consumption  (7000  kg/hr)  divided  by  IHP. 
Three  Cylinder  Locomotive. 
Loss  area  A  -j-  B    kg/sqcm 

7.45 
2.7 

6.47 
0.67 

6.2 
0.33 

Compression  gain  0,28  (A-\-  B)     ....  kg/sqcm 
Mean  effective  pressure,  with  ejector  effect  kg/sqcm 
Gain  due  to  ejector  effect  °/o 

0.75 
6.85 
12.3 

0.19 
4.02 
5.0 

0093 
2.923 
3.0 

Indicated  HP 

940 

1110 

1200 

Steam  consumption  (7000  kg/hr)  divided  by    IHP. 

7.45 

6.3 

5.8 

not  much  more  could  be  desired.  The  saving  of  12%  for  late  cut-offs  is  noteworthy, 
because  it  is  based  on  very  conservative  assumptions.  Furthermore,  the  exhaust  ejector 
effect  allows  of  a  considerable  reduction  of  clearance  volume;  for  instance,  in  the  case 
under  consideration,  from  17  to  11%.  Taking  into  account  the  saving  due  to  the 
una-flow  principle  a  total  saving  of  15%  may  be  expected  with  certainty.  This  saving 
is  all  the  more  important  since  it  occurs  at  heavy  loads  and  therefore  increases  the 
hauling  power  of  the  locomotive  by  this  amount. 

Although  in  describing  the  exhaust  ejector  principle  locomotives  were  con- 
sidered exclusively,  its  field  of  usefulness  is  not  limited  to  the  latter.  It  may  also 
be  found  advantageous  in  street  railway  locomotives,  road  rollers  and  stationary 
engines,  as  well  as  locomobiles  which  are  still  frequently  built  with  two  cylinders 
so  that  they  may  be  started  from  any  crank  position,  which  is  desirable  for  instance 
in  peat  pressing  plants. 


118 


8.  Prof.  Dr.  Nagel's  Experiments. 

A  series  of  extremely  interesting  experiments  on  the  temperature  conditions 
in  a  una-flow  cylinder  were  conducted  by  Prof.  Nagel  in  the  engineering  labora- 
tory of  the  Technische  Hochschule,  Dresden,  and  described  by  him  in  the  Zeit- 


schrift  des  Vereines  deutscher  Ingenieure  Vol.  1913,  No.  27,   July  5.    Part  of  his 
report  is  as  follows: 

"After   Prof.  Stumpf  had   published  his  first  communications  on  the 
character  and  success  of  the  una-flow  steam  engine  a  number  of  years  ago, 


119 


Dr.  Mollier  and  the  author  approached  the  Verein  deutscher  Ingenieure  with 
the  request  for  an  appropriation  for  investigating  the  temperature  conditions 
in  a  una-flow  cylinder.  This  request  was  granted  in  the  most  whole-hearted 
manner.  At  the  same  time  the  Saxon  Government  provided  considerable 
sums  for  the  completion  of  the  testing  plant.  The  una-flow  cylinder  used  for 
this  purpose  in  the  engineering  laboratory  of  the  Technische  Hochschule, 
Dresden,  was  built  by  the  Ntirnberg  Works  of  the  Maschinenfabrik  Augsburg- 
Ntirnberg,  and  took  the  place  of  the  low  pressure  cylinder  of  the  existing  triple- 
expansion  engine.  It  was  put  into  operation  during  September  1911.  The 
cylinder  as  shown  in  Fig.  1,  has  a  bore  of  450  mm  and  a  stroke  of  650  mm, 
and  the  engine  runs  at  150  r.  p.  m. 

In  order  to  determine  the  thermal  peculiarities  of  the  Stumpf  cycle  it 
was  planned  to  measure  the  temperature  changes  of  the  working  steam  at 
different  points  in  the  cylinder.  It  was  later  also  found  desirable  to  measure 


ffo/ben 


OecM 


Fig.  2.     (Kolben  =  piston;  Deckel  =  cover). 

the  temperature  variation  of  the  cylinder  wall.  The  determination  of  the  steam 
temperatures  offered  great  difficulties.  It  was  at  first  attempted  to  use  thermo- 
couples of  copper  and  constantan  wire  of  0.2  mm  diameter.  Tests  of  a  similar 
nature  on  a  counterflow  engine  in  the  laboratory,  using  the  same  elements, 
had  been  started  five  years  ago,  but  a  critical  examination  showed  that  their 
sensitiveness  is  by  no  means  sufficient  to  follow  the  changes  of  temperature 
with  the  necessary  speed.  After  long  and  futile  experiments  with  thermo- 
couples composed  of  thinner  wires  down  to  0.07  mm  diameter,  it  was  found, 
according  to  a  test  report  published  in  an  American  periodical,  that  wires  of 
so  small  a  diameter  as  0.01  mm  were  required  for  the  thermocouple  to  be  suffi- 
ciently sensitive.  It  seemed  impossible  to  produce  thermocouples  with  wires 
of  this  thinness  on  account  of  the  difficulty  in  making  the  junction,  and  for 
this  reason  the  use  of  electric  resistance  thermometers  was  decided  upon  early 
in  1912.  The  material  for  the  latter  was  obtained  in  the  form  of  drawn  tungsten 
filaments  as  used  in  electric  lamps.  A  wire  of  about  50  mm  length  was  wound 
in  zigzags  upon  a  glass  frame  provided  with  platinum  hooks  for  this  purpose, 
as  shown  in  Fig.  2.  The  main  difficulty  was  a  satisfactory  connection  of  the 
resistance  unit  to  the  lead  wires  in  order  to  enable  the  termometers  so.  con- 


120 


structed  to  resist  the  effect  of  the  steam  currents  inside  the  cylinder.  The 
measurement  of  wall  temperatures  was  rendered  difficult  by  the  fact  that  the 
insertion  of  the  measuring  unit  into  the  cylinder  wall  necessitates  the  drilling 
of  a  hole  which  more  or  less  disturbs  the  heat  flow.  This  may  be  the  cause 
of  an  erroneous  temperature  indication.  In  order  to  reduce  this  possibility  to 
the  utmost,  the  arrangement  illustrated  in  Fig.  3  to  5  was  employed.  A  hole 
of  15  mm  diameter  was  drilled  in  the  cylinder  wall,  into  which  was  closely  fitted 
a  cast  iron  plug  having  a  hole  of  9  mm  diameter  bored  to  within  0.5  mm  of 
the  bottom.  Into  this  hole  was  fitted  a  second  cast  iron  plug  having  two  drilled 
holes  of  2  mm  diameter  from  end  to  end.  These  holes  contained  the  copper 
and  constantan  wires  of  0.1  mm  diameter  insulated  by  small  glass  tubes,  their 


Fig.  5. 

ends  being  embedded  in  grooves  at  the  bottom  surface  of  the  cast  iron  plug. 
A  thermocouple  of  similar  construction  was  also  fitted  to  the  piston,  as  indi- 
cated at  ft,  in  Fig.  1,  and  in  Figs.  6  and  7.  The  leads  of  this  thermocouple 
were  carried  through  the  hollow  tail  rod,  provided  with  a  porcelain  lining  for 
this  purpose. 

For  measuring  the  change  in  voltage  corresponding  to  the  changes  in 
temperature,  a  galvanometer  made  by  Edelmann  in  Munich  was  employed. 
According  to  Fig.  8,  it  consists  of  a  powerful  electromagnet  which  is  supplied 
with  current  from  a  storage  battery.  In  the  magnetic  field  is  stretched  a  fila- 
ment of  gold  or  platinum  having  a  diameter  of  from  0.002  to  0.005  mm,  which 
carries  the  current  to  be  measured.  The  displacement  of  this  wire  due  to  electro- 


121 

magnetic  forces  is  a  measure  of  the  current  flowing  through  the  circuit,  and 
therefore  also  a  measure  of  the  temperature.  This  displacement,  although 
amounting  to  only  a  fraction  of  a  millimeter,  is  projected  on  an  enlarged  scale 
onto  the  focal  plane  of  a  camera  by  means  of  a  beam  of  light  from  a  source  L 
and  a  system  of  microscope  lenses.  The  photographic  plate  is  moved  pro- 
portionately to  the  piston  travel  or  crank  angle  behind  a  slit  in  the  focal  plane, 


Fig.  6  u.  7. 

thus  producing  a  photographic  record  of  the  changes  of  temperature  with 
stroke  or  time.  Special  methods  were  devised  for  rapid  calibration  of  the  dis- 
placement of  the  filament.  In  Figs.  9  and  10  are  reproduced  two  records  of 
steam  and  wall  temperature  based  on  time.  The  remarkable  'feature  about 


Bilet- 


eberie 


Fig.  8.    (Bildebene  =  Focal  plane.) 

the  temperature  change  of  the  working  steam  is  the  fact  that  at  the  end  of 
compression  the  latter  attains  temperatures  of  such  a  magnitude  as  were 
hitherto  thought  impossible.  A  terminal  compression  temperature  of  about 
500°  was  observed  when  running  with  saturated  steam  of  10  at.  gage.  The 
wall  temperature  was  measured  at  the  points  a,  6,  c,  d,  e,  /,  g  and  /c,  indicated 
in  Fig.  1.  The  change  of  temperature  at  the  end  of  the  cylinder  barrel  at  point 
b  is  of  considerable  significance.  A  series  of  tests  made  with  constant  cut-off 


122 


Temperature  time  diagram  of  steam  close  to  cylinder  head  surface.    (Point  of  measurement  a.) 


tintauchttqfe  Q 

— — —  1  Umdrehung 


Fig.  9. 


Temperature  time  diagram  of  cylinder  wall  at  a  depth  of  0,5  mm  from  inside  surface. 

(Point  of  measurement  «/.) 


Fig.  10. 


of  10%  and  saturated  as  well  as  superheated  steam  of  different  temperatures 
showed  that  the  highest  mean  temperature  at  this  point  was  reached  when 
operating  with  saturated  steam;  even  superheated  steam  of  350°  did  not  produce 
the  same  high  wall  temperature.  The  sensitiveness  of  the  thermocouples  was 
raised  to  such  a  degree  that  the  passing  of  every  piston  ring  over  a  point  of 
measurement  produced  a  clearly  discernible  wave.  The  moment  of  passage 
of  the  several  rings  over  the  thermocouple  is  clearly  indicated  in  Fig.  10 
by  the  shading  between  the  various  lines;  the  diagram  was  taken  at 
point  d" 

The  above  report  also  includes  descriptions  of  the  several  instruments  which 
were  used  during  the  tests  and  for  the  analysis  of  their  results.  Among  others, 
there  are  mentioned  a  harmonic  analyser  by  Mader,  an  instrument  fitted  with 
a  microscope  for  measuring  indicator  cards,  made  by  H.  Maihak,  of  Hamburg, 
and  an  apparatus  furnished  by  Steinmuller,  of  Gummersbach,  for  automatically 
measuring  the  condensate,  which  proved  to  be  very  exact. 

The  temperature  diagram  in  the  above  report  by  Prof.  Nagel  merits  parti- 
cular attention.  Instead  of  a  terminal  compression  temperature  of  500°  for  3.3% 
clearance,  a  final  temperature  of  900°  should  be  obtainable  with  a  clearance  volume 
of  1%.  While  on  the  one  hand  the  terminal  compression  temperature  was  500° 
for  saturated  steam,  it  decreased  to  480°  or  450°  for  increasing  degrees  of  super- 
heat. This  may  probably  be  attributed  to  the  more  energetic  heating  action  of 
the  steam  jacket  in  the  case  of  saturated  steam.  Prof.  Nagel  further  states  that 
the  temperature  at  the  point  b  of  the  cylinder  wall,  for  12%  cut-off  and  saturated 
steam  of  184°  in  the  jacket,  was  128°,  which  fell  to  111°  for  the  same  cut-off  and 
superheated  steam  of  220°;  and  again  slowly  rose  to  118°  for  a  further  increase 
of  the  initial  steam  temperature  to  350°.  The  temperature  of  the  inner  cylinder 
head  surface  was  177°  for  an  initial  or  jacket  steam  temperature  of  184°,  the  total 
variation  during  one  revolution  being  only  0,5°.  The  temperatures  at  the  points 
*,  c,  d,  (Fig.  1)  were  found  to  be  128°,  102°  and  83°,  with  a  total  fluctuation  of 
3°,  3°  and  2.8°  respectively.  At  the  point  k  on  the  piston,  distant  36  mm  from 
the  cylinder  wall,  the  temperature  was  found  to  be  164.5°  with  a  total  fluctuation 
of  1.3°.  Attention  is  especially  called  to  the  latter  figures,  since  they  prove  the 
statements  previously  made  concerning  the  favorable  thermal  action  of  the  piston 
head  surface.  At  the  points  of  measurement  the  heat  had  to  penetrate  a  metal 
thickness  of  0.65  mm.  It  will  also  be  noticed  in  Fig.  9  that  a  very  pronounced  kink 
occurs  in  the  temperature  curve  where  the  steam  changes  from  the  saturated  to 
the  superheated  state,  and  also  that  an  abrupt  drop  in  temperature  takes  place 
at  the  moment  of  admission,  from  the  high  terminal  compression  temperature  of 
about  530°  to  that  of  the  live  steam. 

The  contrary  effect  of  heating  by  the  jacket  steam  and  cooling  by  the  cylinder 
steam  at  the  point  a  is  also  evident  in  Fig.  9,  as  well  as  the  corresponding  small 
temperature  fluctuation  at  this  point  during  the  complete  cycle,  considered  apart 
from  the  sudden  rise  due  to  the  heat  of  compression.  The  comparatively  high 
mean  temperature  and  small  fluctuation  at  the  point  d  are  also  noticeable  in 
Fig.  10. 


124 


In  Fig.  11  is  shown  an  especially  clear  temperature  diagram  in  which  the 
kinks  in  the  compression  and  expansion  lines  corresponding  to  the  change  from 
saturated  to  superheated  steam  are  clearly  noticeable. 

There  is  a  surprisingly  high  temperature  during  the  last  part  of  expansion, 
the  first  part  of  compression  and  especially  during  exhaust  (about  100°), 
although  the  engine  was  operated  with  a  vacuum  of  98%.  As  this  card  was 
taken  at  the  cylinder  side  of  the  cover,  the  explanation  is  easily  found  in  the 
great  flow  of  heat  from  the  cover  to  the  working  steam  during  that  time. 


3.  C.  Cooer  sieff . 


I0f 


J).C.  Cr 


Fig.  11. 

The  drop  of  temperature  through  the  cylinder  head  wall  was  found  to  be 
7°  to  8°  for  saturated  steam,  15°  for  steam  of  250°,  and  25°  for  steam  of  350°  initial 
or  jacket  temperature. 

A  close  study  of  the  temperature  diagrams  given  in  Figs.  9  and  10  and  11 
has  as  its  final  result  a  confirmation  of  the  thermal  advantages  of  the  una-flow 
principle  and  the  jacketing  of  the  heads. 

This  is  still  more  emphasized  by  the  comparison  of  the  una-flow  temperature 
diagram  (Fig.  11)  taken  by  Prof.  Nagel  with  a  counter- flow  temperature  diagram 
taken  by  E.  T.  Adams  &  T.  Hall  from  a  common  slide  valve  engine  of  the 
Sibley  College-Cornell  University,  as  shown  in  Fig.  12.  The  comparison  elucidates 
the  striking  thermal  difference  between  both  engines.  Whereas  the  una-flow 
engine  shows  the  highest  temperature  at  the  inlet  end  and  the  lowest  at  the 
exhaust  end,  the  counter-flow  engine  shows  quite  a  thermal  mixture  distributed 
over  both  strokes.  Interesting  is  the  postponement  of  the  phases  of  high  metal 
temperature  caused  by  the  preceding  phases  of  high  steam  temperature  in  the 
counter-flow  engine.  - 


125 


126 


II.  1.  The  Una-Flow  Stationary  Engine. 

The  una-flow  engine  has  found  a  very  wide  use  as  a  stationary  prime  mover 
mainly  by  reason  of  its  simplicity,  its  straight  line  construction,  its  high  economy 
and  its  adaptability  to  changing  load  requirements.  The  tandem  counterflow  engine 
which  still  comes  occasionally  into  competition  with  it,  is  at  a  disadvantage  on 
account  of  its  two  cylinders,  its  two  pistons,  its  piping  and  the  inaccessibility 
of  its  exhaust  valves. 

The  conditions  of  close  regulation  required  of  stationary  engines,  in  which 
may  be  included  engines  for  electric  current  generation,  are  satisfied  in  the  una- 
flow  engine  in  the  best  possible  manner  since  the  action  of  the  governor  is  direct 
and  is  not  impeded  by  steam  already  contained  in  the  engine,  as  is  the  case  in 
multiple  expansion  engines  where  the  effect  of  such  steam  on  the  regulation  makes 
itself  unpleasantly  noticeable. 

The  range  of  cut-off  in  una-flow'  engines  is  usually  from  0  to  25%,  although 
cut-offs  are  found  up  to  40%,  or  even  50  or  60%  for  instance  in  rolling  mill  engines. 
The  range  of  the  governor  must  include  zero  cut-off,  in  which  case  the  inlet  valve 
does  not  open  at  all.  The  lead  of  the  steam  valves  at  all  cut-offs  must  be  kept 
down  to  the  minimum  or  reduced  to  nothing  if  possible,  since  large  lead  causes 
condensing  engines  to  knock  badly,  especially  if  the  clearance  is  large  and  the 
vacuum  high.  Non-condensing  engines  with  large  clearance  and  long  compression 
always  run  quietly  and  can  therefore  stand  more  lead. 

It  is  easily  possible  to  start  engines  having  25%  maximum  cut-off  even  under 
load  and  with  a  directly  driven  air  pump,  more  particularly  if  an  additional  clearance 
space  is  provided  and  opened  up  during  the  first  strokes  until  sufficient  vacuum 
is  generated.  The  best  location  for  these  additional  clearance  pockets  is  in  the 
cylinder  head  opposite  the  end  of  the  cylinder,  so  that  for  condensing  service  they 
will  act  as  a  very  effective  insulation  between  cylinder  head  and  frame,  while  pro- 
viding double  the  amount  of  cover  jacket  surface  for  non-condensing  operation. 

Every  condensing  stationary  una-flow  engine  should  be  equipped  with  addi- 
tional clearance  spaces  in  order  to  facilitate  starting  if  the  air  pump  is  directly 
driven,  and  to  allow  of  running  the  engine  without  the  condenser. 

The  clearance  volume  averages  about  1,5  to  2%  for  condensing  operation 
and  high  vacuum,  and  13  to  28.%  if  the  additional  clearance  spaces  are  opened  up 
for  non-condensing  service  (see  page  48). 

In  non-condensing  engines  the  necessary  clearance  may  be  arranged  in  the 
cupped  ends  of  the  piston.  On  account  of  the  large  work  of  compression  such 
engines  require  comparatively  heavy  flywheels. 

Since  the  una-flow  engine  has  only  two  inlet  valves,  the  use  of  a  lay-shaft 
is  unnecessary.  As  is  shown  in  Figs.  1  to  3  of  this  chapter,  and  in  Figs.  4  and  5 
(page  10),  the  inlet  valves  may  be  driven  from  an  eccentric  on  the  crankshaft 


127 


W/////////////////////////////////'/'///"' 

Fig.  1. 


Fig.  2. 


Fig.  3. 


128 

acted  on  by  a  shaft  governor,  by  means  of  a  rocker  arm  and  cam  mechanism  (Stumpf 
gear).  A  lay-shaft  with  its  bevel  gears  and  bearings  is  thus  dispensed  with.  Con- 
sideration must  be  given  to  the  expansion  of  the  cylinder.  If  the  latter  is  pro- 
vided with  steam  jackets  receiving  their  supply  from  a  connection  to  the  steam 
pipe  ahead  of  the  main  stop  valve,  then  the  cylinder  may  be  warmed  up  prior  to 
starting,  and  the  valve  gear,  if  set  correctly  for  the  hot  engine,  will  give  proper 
distribution  from  the  very  start.  If  the  cylinder  is  unjacketed,  then  the  steam 
distribution  will  be  incorrect  for  a  while  after  starting  until  the  cylinder  has  reached 
its  expanded  condition.  It  is  always  advisable  to  design  the  valve  gear  with  out- 
side admission,  or  in  other  words  to  arrange  cams  and  rollers  so  as  to  make  their 
action  conform  to  the  steam  lap  of  a  slide  valve,  so  that  while  the  cylinder  is  still 
insufficiently  heated,  the  head  end  valve  will  open  late  instead  of  too  early.  This 
negative  lead  combined  with  a  simultaneous  earlier  cut-off  will  do  less  harm  than 
an  early  opening  of  the  valve,  which  may  cause  the  engine  to  knock. 


Fig.  4. 


There  remains  another  possibility  of  correcting  the  bad  influence  of  the  expan- 
sion of  the  cylinder  even  if  the  latter  is  not  provided  with  steam  jackets,  by  ex- 
tending the  cylinder  lagging  around  the  rod  between  the  valve  bonnets  (Fig.  4), 
thus  heating  it  to  approximately  the  mean  cylinder  temperature.  It  is  also  possible  to 
drive  the  head  end  valve  through  an  equal-armed  rocker  mounted  at  the  center  line 
of  the  exhaust  belt.  This  insures  permanent  correct  motion  for  the  head  end  valve. 

If  a  lay-shaft  is  used,  the  influence  of  the  cylinder  expansion  is  eliminated, 
and  the  steam  distribution  must  always  be  correct. 

Fig.  5  illustrates  details  of  a  valve  bonnet  as  used  with  the  Stumpf  gear. 
The  cam  is  connected  to  the  valve  crosshead,  and  the  reciprocating  slide  is  grooved 
to  accommodate  the  roller  and  at  the  same  time  form  an  oil  bath.  The  guide  for 
the  reciprocating  slide  is  long  enough  so  that  the  groove  never  runs  beyond  it, 
the  loss  of  oil  by  splashing  and  the  entrance  of  dust  thus  being  prevented.  The 
oil  collecting  in  the  groove  is  transferred  by  the  roller  to  the  cam,  so  that  perfect 
lubrication  and  reliable  operation  of  these  important  parts  is  insured. 


129 


Fig.  6  shows  a  twin  una-flow  engine  with  Stumpf  gear,  in  which  a  jack  shaft 
having  two  crank  throws  is  driven  by  a  pair  of  eccentrics  on  the  crank  shaft  set 
at  90°.  This  jack  shaft  carries  a  shaft  governor  acting  upon  an  eccentric  on  each 
side  of  it,  which  operates  the  valve  mechanism  of  its  corresponding  cylinder  through 
a  rocker  arm.  In  this  way  the  valve  gears  are  positively  connected,  so  that  both 
of  them  always  give  the  same  cut-off.  The  short  vertical  eccentric  rod  also  helps 
to  equalize  the  cut-offs  of  both  cylinder  ends,  and  the  small  diameter  of  the  jack 
shaft  facilitates  the  design  of  the  governor.  This  engine  possesses  great  reserve 
power  since  each  half  is  able  to  carry  the  whole  load. 

The  elimination  of  exhaust  valves  and 
their  gear  will  be  found  very  convenient  in 
horizontal  engines,  since  it  leaves  the  whole 
space  underneath  the  cylinder  free  for 
piping  and  permits  of  a  close  arrangement 
of  the  condenser.  (See  Figs.  2  to  5,  chapter 
I,  3b,  p.  69—70.) 

The  Erste  B runner  Maschinenfabrik- 
gesellschaft  was  the  first  concern  to  take 
up  the  una-flow  engine,  and  decided  to  re- 
build an  old  80  HP.  single  cylinder  con- 
densing engine  with  a  forked  frame  by 
fitting  it  with  a  una-flow  cylinder,  designed 
by  the  author,  having  a  bore  of  400  mm 
and  a  stroke  of  420  mm  (Fig.  7).  On  the 
free  end  of  the  crank  shaft  was  mounted  a 
shaft  governor  acting  on  the  eccentric  ope- 
rating the  inlet  valves  by  means  of  a  rocker 
arm  on  the  exhaust  belt  and  a  pair  of 
Lentz  cam  mechanisms.  Although  this  first 
design  was  susceptible  of  improvement  in 
many  respects,  its  economy  even  with 
rather  low  vacuum  was  'equal  to  that  of 
a  compound  engine  of  the  same  size. 

Fig.  8  shows  another  engine  built  by 
the  same  Company. 

Shortly  after  the  latter  had  taken  up 
this  work,  the  Elsassische  Maschinenfabrik 
decided  on  a  large  scale  experiment.  Their 
first  una-flow  engine,  built  to  the  author's 

design,  had  a  cylinder  bore  of  640  mm  and  a  stroke  of  1000  mm,  with  a 
rated  load  of  500  HP.  (Figs.  1  and  9).  This  engine,  which  was  directly  con- 
nected to  an  electric  generator,  was  tested  by  the  Elsassische  Verein  der  Dampf- 
kessel-Besitzer  (Alsatian  Association  of  Steam  Boiler  Owners),  on  February  21, 
1909.  The  result  of  a  trial  of  four  hours  and  eight  minutes  duration  showed  a  steam 
consumption  of  4.6  kg/I  HP-hour  for  an  initial  steam  pressure  of  12.6  at.  gage 
and  a  temperature  of  331°  G,  at  a  speed  of  121  r.  p.  m.  This  is  a  very  creditable 

Stumpf,  The  una-flow  steam  engine.  9 


Fig.  5. 


130 


131 

result  if  it  is  borne  in  mind  that  the  engine  did  not  derive  the  full  benefit  from 
the  vacuum  on  account  of  too  small  an  exhaust  pipe  and  the  use  of  an  oil  separator 
between  cylinder  and  condenser.  (Back  pressure  0,145  at.  abs.)  By  correct  design 


Fig.  7. 

of  the  condensing  equipment  in  the  way  previously  suggested,  by  jacketing  of  the 
cylinder,  and  the  use  of  tighter  valves,  the  steam  consumption  could  be  conside- 
rably diminished,  as  proved  by  later  engines  built  by  the  same  makers. 


Fig.  8. 


Figs.  2  and  3  show  a  una-flow  engine  of  900  HP  rated  load  built  by  the  same 
firm.    Noteworthy  is  the  heavy  frame,  the  engine  being  of  center  crank  construc- 


9* 


132 

tion  which  is  now  used  by  several  concerns.  The  center  crank  type  is  advantageous 
where  the  forces  on  the  moving  parts  are  heavy,  especially  for  large  short  stroke 
engines. 

In  Figs.  10  and  11  is  shown  still  another  engine  by  the  same  makers,  as  well 
as  its  governor  and  valve  gear  parts. 


Fig.  9. 

A  una-flow  engine  built  by  Burmeister  &  Wain,  of  Copenhagen,  and  designed 
by  the  author,  is  shown  in  Figs.  12  and  13.  The  direct  connection  of  the  con- 
denser to  the  cylinder  should  be  noted,  as  well  as  the  method  of  supporting  the 


Fig.  10. 

rear  end  of  the  latter  on  two  adjustable  rods,  and  the  simple  air  pump  drive.  The 
cylinder  is  left  unjacketed  on  account  of  the  use  of  superheated  steam  (Fig.  14). 
The  head  jackets,  however,  are  carried  up  to  the  point  of  normal  cut-off.  The 


133 


Fig.  11. 


Fie.  12. 


134 

piston  is  turned  to  a  smaller  diameter  for  a  corresponding  distance  to  provide  for 
expansion.  The  additional  clearance  pockets  are  fitted  with  two  clearance  valves, 
one  one  each  side  of  the  inlet,  one  of  which  is  sufficient  for  starting,  while  the 


Fig.  14. 


second  one  has  to  be  opened  for  non-condensing  operation  at  full  speed  (Fig.  15). 
The  area  of  contact  between  cylinder  head  and  frame  is  kept  as  small  as  possible 
in  order  to  reduce  the  conduction  of  heat  to  a  minimum.  The  cylinder  head 
casting  is  perfectly  symmetrical  so  as  to  increase  its  range  of  usefulness. 


135 


Fig.  15. 


Fig.  16. 


136 


Fig.  17. 


Fig.  18. 


137 

Tests  made    with  this  engine    by   Mr.  Bacher,   Professor   at    the  Technical 
Hochschule  in  Copenhagen,  showed  the  following  results. 


Load 

Pressure 

kg/sqcm 

Steam 
Tempe- 
rature 

°C 

Vacuum 
in  •/.  of 
760  mm 

r.  p.m. 

Steam 
Con- 
sumption 
kg/hr. 

Steam  Consumption 

KW 

BHP 

IHP 

KW/nr. 

BHP/hr. 

IHP/hr. 

64.50 

98.7 

116.0 

9.90 

352 

94.0 

179.0 

479.0 

7.44 

4.86 

4.12 

86.46 

130.0 

149.0 

9.87 

354 

93.8 

175.0 

632.0 

7.32 

4.86 

4.24 

108.66 

163.0 

184.5 

9.84 

353 

93.5 

176.5 

798.4 

7.36 

4.90 

4.34 

131.24 

197.0 

222.0 

9.80 

353 

92.6 

173.5 

976.7 

7.45 

4.97 

4.40 

109.00 

164.0 

186.0 

9.75 

dry 
saturated 

93.0 

178.0 

1150.6 

10.55 

7.03 

6.20 

In  view  of  the  omission  of  the  cylinder  jackets,  these  results  are  in  close 
agreement  with  the  tests  of  a  300  HP  una-flow  engine  given  on  p.  11. 

A  single-acting  una-flow  engine  built  by  Burmeister  &  Wain  is  shown  in 
Fig.  16.  Engines  of  this  type  are  widely  used  in  Danish  dairies.  The  horizontal 
valve  is  operated  by  a  cam  mechanism  directly  connected  to  a  shifting  eccentric 
on  the  crank  shaft.  (Figs.  17  and  18.) 

A  stationary  engine  built  by  Ehrhardt  &  Sehmer,  of  Saarbriicken,  for  the 
power  plant  of  the  Saar  Valley  Railway,  is  shown  in  Fig.  19.  This  engine  has 


Fi?.  19. 


a  cylinder  bore  of  650  mm,  a  stroke  of  1000  mm,  and  a  rated  load  of  500  HP  at 
a  speed  of  130  r.  p.  m.  It  has  run  for  long  periods  at  an  overload  of  nearly  100%. 
The  average  cut-off  of  una-flow  engines  at  rated  load  being  only  about  10%,  their 


138 


139 


capacity  for  overload  is 
far  greater  than  that  of 
any  other  type  of  engine. 
If  the  dimensions  of  the 
driving  parts  are  based 
upon  the  initial  pressure 
less  the  inertia,  then  even 
a  heavy  overload  does 
not  materially  increase 
the  load  upon  them. 

In  Fig.  20  is  shown 
the  largest  una-flow  en- 
gine so  far  built,  con- 
structed by  Ehrhardt  & 
Sehmer  for  driving  a  rol- 
ling mill  at  the  steel 
works  of  Gebrtider  Roch- 
ling,  atVolklingen,  having 
acylinderboreoflVOOmm 
and  a  stroke  of  1400  mm, 
the  speed  being  110  to 
130  r.  p.  m.  The  cylinder 
was  made  with  this  large 
bore  on  account  of  the 
low  steam  pressure  avail- 
able at  the  time,  and  is 
provided  with  two  inlet 
valves  at  each  end  for 
this  reason.  Later  on, 
when  new  high  pressure 
boilers  have  been  instal 
led,  it  is  intended  to 
substitute  for  the  present 
cylinder  a  smaller  one 
with  only  one  valve  at 
each  end. 

Another  una  -  flow 
engine  built  by  the  same 
firm  and  delivered  to  the 
Aplerbeck  steel  works, 
is  illustrated  in  Figs.  21 
and  22.  The  cylinder  bore 
is  1450  mm,  the  stroke 
1500  mm,  the  speed  100 
r.  p.  m.,  and  the  steam 
pressure  8  at.  gage. 


140 


141 


142 


143 

This  engine  has  nearly  the  same  dimensions  of  driving  parts  as  the  one  just  de- 
scribed. The  diameter  of  the  piston  rod  is  250  mm,  that  of  the  tail  rod  225  mm, 
the  crosshead  pin  is  400  mm  diameter  by  600  mm  long,  the  crank  pin  550  mm 
diameter  by  600  mm  long,  and  the  main  bearing  730  mm  diameter  by  1200  mm 
long.  The  use  of  a  side  crank  in  such  a  large  engine  of  short  stroke  is  noteworthy. 
In  order  to  reduce  the  overhang,  the  crank  and  crank  pin  are  of  cast  steel  in  one 
piece,  with  a  hub  length  of  only  450  mm  (hub  length  :  shaft  =  0.62).  The  side 
crank  construction,  together  with  its  corresponding  type  of  frame,  makes  the  engine 
simple  and  inexpensive.  The  same  cannot  perhaps  be  said  of  the  Zvonicek  valve 
gear  employed  on  this  engine,  but  it  has  the  advantage  of  giving  the  late  cut- 
offs essential  for  rolling  mill  engines,  and  of  permitting  the  use  of  a  standard  governor 
which  is  in  many  cases  preferred  to  a  shaft  governor  on  account  of  its  simplicity 
and  accessibility.  The  Zvonicek  gear  consists  of  a  fixed  eccentric,  the  strap  of  which 


Fig.  25. 


is  provided  with  a  cam  profile  and  held  at  its  armlike  extension  under  the  control  of 
the  governor.  The  combined  motion  of  eccentric  and  cam  is  transmitted  to  the 
valve  bonnet  cam  mechanism  by  a  reach  rod  provided  with  a  roller  at  its  lower  end. 

Figs.  23  and  24  illustrate  clearly  the  trend  of  development  due  to  the  una- 
flow  engine  and  show  the  replacement  of  the  two  cylinders  of  an  old  tandem 
counterflow  engine  by  a  single  una-flow  cylinder.  A  number  of  such  reconstruc- 
tions have  been  carried  out  by  Ehrhardt  &  Sehmer  and  other  firms. 

An  engine  built  by  Musgrave  &  Sons,  Ltd.  of  Globe  Iron  Works,  Bolton, 
England,  is  shown  in  Fig.  25.  This  firm  is  credited  with  the  introduction  of  the 
una-flow  engine  on  a  large  scale  in  Great  Britain  and  colonies,  the  Stumpf  valve 
gear  being  employed  exclusively.  A  test  carried  out  by  Mr.  F.  Thomas  on  one 
of  their  engines,  having  a  cylinder  bore  of  685.8  mm  and  a  stroke  of  914.4  mm 
gave  the  following  results :  Steam  pressure  10.67  at.  gage  at  the  throttle,  superheat 
10°,  speed  129  r.  p.  m.,  vacuum  at  the  cylinder  66  cm,  load  317  I  HP,  and  steam 
consumption  4,98  kg/I  HP-hour.  The  cylinder  barrel  was  unjacketed. 


144 

Stork  &  Co.,  of  Hengelo,  Holland,  also  employ  only  the  Stumpf  gear  on  their 
engines,  one  of  which  is  shown  in  Fig.  26.  This  firm  has  been  very  successful  in  in- 
troducing the  una-flow  engine  in  Holland  and  the  Dutch  colonies.  Stork  &  Go. 
report  that  a  test  of  one  of  their  engines  (650  mm  bore  by  900  mm  stroke,  speed 
125  r.  p.  m.)  showed  a  steam  consumption  of  4.86  kg/I  HP-hour,  the  steam  pressure 
being  8.24  at.  gage,  and  the  temperature  248°  G.  The  cylinder  barrel  was  un jacketed. 

The  type  of  una-flow  engine  built  by  the  Maschinenfabrik  Augsburg- Niirnberg 
is  shown  in  Figs.  27  and  28.  The  inlet  valves  are  placed  horizontally  and  are  ope- 
rated by  a  rocking  shaft  and  cam  mechanism.  This  cam  receives  its  motion  from 
a  short  jack  shaft  which  in  turn  is  driven  by  the  governor  eccentric  on  the  lay 
shaft.  The  clearance  valve  is  situated  opposite  the  steam  valve  and  is  arranged 
to  act  automatically  in  case  of  sudden  failure  of  the  vacuum.  Partially  or  entirely 
unbalanced  inlet  and  spring-loaded  clearance  valves  may  be  made  to  serve  the 
same  purpose. 

One  engine  of  this  type  (903  mm  cylinder  bore,  1000  mm  stroke)  was  fur- 
nished to  J.  P.  Stieber,  at  Roth  near  Niirnberg  and  was  tested  by  the  Bayerische 
Revisionsverein  on  February  23,  1912. 

Duration  of  test hours  4.17  4.05  8.06 

Boiler  pressure at.  gage  13.3  13.3  13.2 

Steam  temp,  at  engine 255  291  310 

Steam  pressure  in  cylinder at.  gage  9.3  10.4  11».4 

Vacuum  in  cylinder %  90  90  90 

Vacuum  in  condenser %  93  92  92 

Actual  cut-off    .  ~ %  3  6  12 

Speed .     r.  p.  m.  125.5  123.3  126.3 

Indicated  horse  power HP  473  793  1109 

Steam  consumption  in  kg/I  HP-hour  including  con- 

densate  from  steam  pipe 4.69  4.74  4.71 

Heat  consumption  in  Gal/I HP-hour  based  on  total 

heat  of  steam  entering  the  engine 3320  3440  3460 

Thermal  efficiency %  19  18.4  18.3 

This  engine  was  designed  for  a  steam  temperature  of  330°  G  for  which  reason  the 
cylinder  barrel  was  left  unjacketed.  With  jackets  the  steam  consumption  would  in 
this  case  have  been  considerably  lower  on  account  of  the  beneficial  effect  of  cylinder 
jackets  for  small  cut-offs  and  low  initial  temperatures.  On  the  other  hand,  the 
results  once  more  demonstrate  the  small  variation  in  steam  consumption  for  large 
ranges  of  load  (473  to  1109  HP)  when  no  cylinder  jackets  are  employed. 

An  engine  built  by  the  Gorlitzer  Maschinenbauanstalt  is  shown  in  Fig.  29. 
The  cylinder  has  a  bore  of  1100  mm,  a  stroke  of  1300  mm,  and  the  engine  runs 
at  91.6  r.  p.  m.  The  valve  gear  comprises  a  lay  shaft  with  governor  and  shifting 
eccentric  acting  on  short  rocking  shafts  alongside  the  cylinder.  The  ends  of  the 
cylinder  are  jacketed.  The  air  pump  is  driven  from  an  extension  of  the  tail  rod. 

Fig.  30  shows  one  of  the  latest  engines  built  by  this  Company.  There  is  only 
one  governor  eccentric,  and  the  valves  are  operated  through  Lentz  cam  mechanisms 
from  a  rocking  shaft  on  the  cylinder,  having  two  levers  set  at  180°.  The  air  pump 


145 


bb 

s 


Stumpf,  The  una-flovv  steam  engine. 


10 


146 


tc 

£ 


147 


10* 


149 


150 

is  again  driven  from  the  tail  rod.  In  the  single-eccentric  type  of  valve  gear  the 
lay-shaft  as  well  as  the  hole  in  the  latter  for  the  synchronizing  device  are  shorter, 
and  the  governor  may  be  placed  close  to  the  rear  bearing.  The  basic  idea  of  this 
gear  is  similar  to  the  one  used  by  the  Maschinenfabrik  Augsburg-  Nurnberg.  The 
single-eccentric  gear  is  also  described  in  the  Z.  d.  V.  d.  I.,  1914,  No.  19,  page  729. 
The  valves  are  placed  on  the  cylinder  and  consequently  have  somewhat  larger 
clearance  volume  and  surfaces.  The  jacketing  is  excellent;  the  arrangement  of 
the  condenser,  however,  is  not  free  from  objections.  The  firm  reports  the  following 
steam  consumption  results. 


Mean 

Engine 

Rated 
Load 

Stroke 

Bore 

Steam 
Pressure 
at.  gage 

Steam 
Temp. 

r.  p.'m. 

Steam 
Consumption 
kg/IHP-hr. 

Measurements 
based  on" 

HP 

mm 

mm 

°C 

1 

550 

1000 

750 

11.8 

250 

127.2 

4.8 

1 

2 

810 

1300 

850 

9.5 

269.7 

91.6  - 

4.99 

Boiler 

3 

300 

800 

650 

11.7 

253 

15.67 

488 

[       feed  water 

4 

523 

800 

600 

11.5 

231 

152 

4.88 

J 

5 

140    |      600 

1 

375 

9.7 

340 

200 

4.38 

Condensate 

The  steam  temperatures  in  engines  1  and  5  were  measured  at  the  entrance 
to  the  cylinder  head,  and  in  engines  2,  3  and  4  in  the  middle  of  the  same. 


Fig.  31. 

The  greatest  credit  for  the  commercial  introduction  of  the  una-flow  engine 
is  due  to  Sulzer  Bros.,  of  Winterthur  and  Ludwigshafen.  The  cylinder  of  their 
first  engine,  which  is^  in  operation  in  the  brass  rolling  mill  of  Wieland  Bros.,  at 
Ulm,  was  designed  by  the  author  (Fig.  4,  chap  I,  Ib,  p.  10).  The  design  of  later 
engines  was  based  on  this  first  one,  the  only  change  being  the  substitution  for 
the  Stumpf  gear  of  a  lay-shaft  gear,  having  two  eccentrics  which  operate  the 
valves  by  means  of  cams  and  roller  levers  pivoted  in  the  valve  bonnets  (Fig.  31). 
The  reciprocating  roller  slide  of  the  Stumpf  gear  has  therefore  been  replaced  by 
the  pivoted  roller  lever.  The  governor  is  placed  close  to  the  rear  lay-shaft  bearing 


in  order  to  reduce  the 
deflection  of  the  shaft 
and  also  to  shorten 
the  bore  required  for 
the  synchronizing  de- 
vice. All  of  the  driving 
parts,  including  those 
of  the  air  pump  in  the 
basement,  are  comple- 
tely enclosed.  A  gear 
pump  on  the  lay-shaft 
supplies  oil  under  a 
pressure  of  about  1  at. 
to  the  bearings  of 
engine  and  air  pump. 
The  oil  collects  in  a 
reservoir  in  the  base- 
ment, where  it  is  fil- 
tered and  again  enters 
the  circulating  system. 
Low  oil  consumption 
and  smooth  running 
due  to  the  oil  cushion 
in  the  bearings  are  ad- 
vantages of  this  sy- 
stem. The  consump- 
tion of  cylinder  oil  is 
also  low  since  only  cne 
cylinder  and  generally 
only  one  piston  rod 
packing  have  to  be 
lubricated,  as  against 
two  cylinders  and  se- 
veral packings  in  an 
ordinary  multi  -  stage 
engine.  The  low  oil 
consumption  is  also 
proved  by  the  high 
mechanical  efficiency. 
An  engine  furnished 
by  Sulzer  Bros,  to  the 
firm  of  Junker  &  Ruh, 
of  Karlsruhe,  having 
a  cylinder  bore  of 
675  mm,  a  stroke  of 
800  mm,  and  a  speed 


152 


of/150  r.  p.  m.,  showed  that  for  a  steam  pressure  of  11  to  12  at.  gage  and  a  tem- 
perature of  250°,  12  to  18  kg  of  cylinder  oil  were  consumed  weekly,  and  8  to  10  kg 
of  bearing  oil  were  added  to  the  circulation  when  running  10  hours  daily  and  six 


CO 

tB 


co 

be 


days  per  week.  The  whole  of  the  oil  in  circulation,  amounting  to  about  1  barrel, 
is  replaced  every  6  to  9  months.  A  hand  pump  is  provided  to  supply  the  bearings 
with  oil  before  starting.  As  shown  in  the  sectional  drawing  Fig.  32,  the  cylinder 
design  incorporates  all  the  essentials  previously  mentioned.  It  is,  however,  to  be 


153 


60 


154 


CO 

bo 


155 

regretted  that  in  many  cases  cylinder  jackets  are  omitted  when  their  use  should 
be  'dictated  by  low  steam  temperatures. 

.  Most  of  the  engines  built  by  Sulzer  Bros,  are  fitted  with  a  valve  design  the 
purpose  of  which  could  be  otherwise  accomplished  simpler  and  better. 

Extensive  experiments  have  enabled  Sulzer  Bros,  to  find  the  proper  mixtures 
for  cylinder  and  piston  castings,  whereby  reliable  operation  of  these  parts  is  in- 
sured without  the  use  of  a  tail  rod.  The  cylinders  are  bored  barrel-shaped  so  that 
the  cylinder  surface  becomes  almost  exactly  cylindrical  under  operating  conditions. 
To  these  precautions,  in  combination  with  a  thoroughly  reliable  lubricating  system, 
must  be  ascribed  the  fact  that  Sulzer  Bros,  have  never  had  piston  troubles.  All 
of  their  una-flow  engines  have  therefore  been  built  with  self-supporting  pistons, 
except  the  engine  shown  in  Figs.  33  and  34,  supplied  to  the  Crefeld  Cotton 
Spinning  Mill,  and  a  series  of  engines  supplied  to  the  Badische  Anilin-  and 
Soda-Fabrik,  where  a  tail  rod  was  used  to  meet  the  purchaser's  wishes. 

A  Sulzer  stationary  engine  of  standard  design  is  shown  in  Fig.  35  (350  BHP 
at  150  r.  p.  m.),  while  Fig.  36  shows  two  Sulzer  una-flow  engines  of  450  BHP 
each,  supplied  to  the  Hafod  Copper  Works,  Swansea,  South  Wales. 

The  Maschinenfabrik  Esslingen  employs  a  particularly  effective  method  of 
boring  una-flow  cylinders  under  temperature  conditions  closely  approaching  those 
of  actual  operation.  The  cylinder  ends  are  heated  to  a  high  temperature  by  ad- 
mitting live  steam  to  the  jackets,  and  the  middle  is  cooled  approximately  to  con- 
denser temperature  by  a  blast  of  air  through  the  exhaust  belt.  The  cylinder  is 
then  bored  cylindrically,  and  the  piston  is  turned  smaller  than  the  cylinder  bore 
with  a  correct  allowance,  a  difference  of  four  thousandths  of  the  diameter  being 
usually  sufficient.  Since  the  piston  expands  more  than  the  cylinder,  and  is  of 
great  length,  ample  bearing  surface  will  be  obtained.  The  piston  heads  should 
be  turned  somewhat  smaller  in  order  to  allow  for  their  greater  expansion. 

A  piston  as  shown  in  Fig.  5,  chap.  I,  4,  p.  77,  fitted  with  bronze  shoes,  offers 
still  greater  safety  against  seizing,  and  this  is  true  to  a  still  greater  degree  of  the 
floating  piston  having  clearance  all  around. 

Piston  troubles  are  caused  in  many  cases  by  a  wrong  method  of  supplying 
oil  to  the  cylinder.  The  force  pump  should  be  timed  in  such  a  manner  that  deli- 
very takes  place  only  while  the  feed  orifice  in  the  cylinder  is  covered  by  the  piston. 
A  good  distribution  of  oil  to  the  piston  and  cylinder  wall  will  then  be  obtained. 
The  oil  feeds  should  preferably  be  placed  in  the  center  of  or  close  to  the  exhaust 
belt,  where  the  cylinder  has  the  lowest  temperature.  One  feed  should  be  arranged 
on  the  vertical  center  line  and  one  each  at  either  side  in  or  below  the  horizontal 
plane,  each  feed  being  supplied  by  a  separate  plunger.  Complaints  which  are 
sometimes  made  regarding  the  high  oil  consumption  of  una-flow  engines  frequently 
arise  from  defective  methods  of  introducing  the  oil.  It  is  fundamentally  wrong 
to  supply  two  or  more  feeds  from  the  same  pump  plunger. 

A  neat  arrangement  of  piping  is  obtainable  if  the  steam  pipes  are  placed  on 
one  side  of  the  engine  and  the  exhaust  pipe,  air  pump,  and  the  cooling  water  and 
discharge  pipes  on  the  other. 

The  following  is  a  report  of  economy  tests  on  the  una-flow  engine  of  the  Cre- 
feld Cotton  Spinning  Mill,  built  by  Sulzer  Bros. 


156 

Steam  was  generated  by  four  Lancashire  (twin  furnace)  boilers,  having  a  total 
heating  surface  of  400  sqm.  A  fifth  boiler,  the  steam  and  feed  lines  of  which  were 
blanked  off  from  the  others,  supplied  steam  for  heating  purposes.  The  feed  water 


Fig.  38. 


was  weighed,  transferred  to  a  large  tank  and  fed  to  the  boilers  by  means  of  a  cen- 
trifugal pump.  Indicator  cards  were  taken  every  10  minutes,  and  the  steam  pres- 
sure, superheat  and  vacuum  were  recorded  at  the  same  intervals.  The  guarantees 
given  for  this  engine  were: 


157 


Maximum  load  2340  I  HP: 

10%  cut-off,     300°  temperature  1590  I  HP  4.45  kg/IHP-hr 
10%        „         350%          „  1530     ,,     4.15 

13%        „         300°  „  1920     „     4.65 

13%        „         350°  „  1860     „     4.35 

These  guarantees  applied  to  steam  of  11.5  at.  gage  pressure  and  condenser 
cooling  water  of  15°  C. 

The  duration  of  the 
test  was  from  8  :  32  A.  M. 
to  4  :  06  P.  M.  a  total  of 
454  minutes. 

The  load  varied  from 
1477  to  1752  I  HP  with  an 
average  of  1632.8  I  HP  at 
a  speed  of  109.5  r.  p.  m. 
Steam  pressure  at  the  cylin- 
der was  11.1  to  12.1,  average 
11.6  at.  gage;  steam  tempe- 
rature at  cylinder  was  260 
to  300°,  average  282.6°  G; 
vacuum  was  71.3  to  73, 1cm, 
average  72.2  cm.  Barometer 
reading  76.4  cm. 

The  temperature  of  the 
cooling  water  was  11.5° 
and  that  of  the  air  pump 
discharge  30.1°  G. 

Since  the  load  was  higher  than  the  guaranteed  figure  of  1590  HP,  and  the 
steam  temperature  less  than  300°,  a  corresponding  correction  of  the  test  results 
was  necessary. 

According  to  the  guarantees,  the  steam  consumption  increases  from  4.45  kg 
to  4.65  kg  or  0.2  kg  for  an  increase  of  load  from  1590  to  1920  HP,  or  330  HP; 
therefore  for  an  increase  in  load  of  1632.8  —  1590  =42.8  HP,  the  permissible 

0  2  •  42  8 

increase  may  be  -— — —  -  =  0.026  kg  and  the  steam  consumption  may  be  4.45 
ooU 

-f-  0.026  —  4.476  kg,  and  still  be  within  the  guarantee. 

Tests  have  shown  that  a  reduction  of  the  initial  temperature  from  300°  to 
282.6°  produces  an  increase  in  steam  consumption  of  3%. 

The  permissible  steam  consumption  may  therefore  be  4.476  x  1.03  =  4.61  kg, 
and  yet  be  within  the  guarantee. 

The  total  steam  consumption  was  56410  kg,  or 


56410-60 


or 


454 

7455.06 
1632.8 


—  7455.06  kg/hour 
==  4.56  kg/I  HP-hour. 


158 


159 

The  guaranteed  figure  was  therefore  satisfied  without  taking  advantage  of 
the  permitted  allowance  of  5%. 

The  above  report  was  made  by  the  Association  for  the  Inspection  of  Steam 
Boilers,  of  Miinchen-Gladbach,  Crefeld  Branch  Office,  on  October  4,  1913,  and 
signed  by  Mr.  Rhenius. 

In  examining  thik  result  it  must  be  borne  in  mind  that  the  cylinder  barrel  of 
this  engine  was  not  jacketed. 

A  una-flow  engine  designed  by  the  author  for  the  Soumy  Machine  Works 
is  shown  in  Figs.  37  to  39.  It  has  valve  gear  of  the  Stumpf  type  and  is  designed 
to  be  used  with  saturated  steam  of  7  at.  gage.  The  cylinder  has  a  bore  and  stroke 
of  450  and  600  mm  respectively,  and  the  speed  is  150  r.  p.  m.  The  resilient  inlet 


Fig.  41. 

valves  and  the  clearance  valves  are  designed  and  arranged  in  such  a  manner  that 
the  total  clearance  volume  amounts  to  only  1.24%  for  a  linear  piston  clearance 
of  3  mm.  The  nut  is  flush  with  the  piston  so  as  to  avoid  the  clearance  volume 
of  about  0.5%  resulting  from  a  projecting  nut.  The  two-piece  cast  steel  self-sup- 
porting piston  is  fitted  with  a  bronze  shoe  fastened  to  it  with  copper  rivets;  the 
rest  of  the  piston  has  several  millimeters  clearance  all  over.  Each  half  carries  three 
somewhat  narrow  rings,  none  of  which  overruns  the  cylinder  bore.  The  harmful 
surfaces  are  small,  and  are  jacketed  and  machined  in  addition.  The  ends  of  the 
cylinder  are  provided  with  jackets  since  saturated  steam  is  used.  The  suction 
of  the  air  pump  takes  place  through  ports,  and  the  discharge  valves  are  arranged 
in  the  heads,  so  that  the  clearance  is  small  and  the  suction  effect  a  maximum. 
The  condenser  is  placed  immediately  under  the  cylinder  with  a  connection  of 
large  area. 

The  cylinder  jackets  are  supplied  with  steam  through  a  separate  pipe  con- 
nected to  the  steam  main  ahead  of  the  stop  valve.    This  allows  the  cylinder  to 


160 


be  warmed  up  before  starting,  and  the  valve  gear  therefore  gives  correct  distri- 
bution from  the  very  beginning.  The  eccentric  rod  is  shortened  and  guided  by 
a  swinging  link  interposed  in  the  valve  gear,  thus  compensating  the  angularity 
of  the  connecting  rod.  This  equalization  of  cut-offs  and  valve  lifts  at  the  two 
cylinder  ends  is  very  complete  for  all  cut-offs,  which  range  from  0  to  25%. 

Fig.  40  shows  another  similar  engine  with  Stumpf  gear  designed  by  the  author 
for  a  company  in  Finland.    The  lower  seats,  instead  of  the  valves,  are  resilient, 

and  the  piston  is  fitted  with  Allan 
metal  rings  to  prevent  seizing.  The 
clearance  volume  is  1,25%. 

Details  of  una-flow  engines 
built  by  the  Ames  Iron  Works, 
of  Oswego,  N.  Y.,  are  given  in 
Figs.  41  to  44.  The  head  and  the 
cylinder  ends  are  jacketed  and  the 
additional  clearance  spaces  are 
arranged  in  the  cylinder  heads. 
The  upper  resilient  seat  of  the 
valves  is  made  of  steel  and  shrunk 
in  place  on  the  cast  iron  valve 
body.  For  non-condensing  service 
the  engine  is  fitted  with  a  piston 
having  cupped  ends  and  the  length 
of  compression  is  90%. 

The  following  test  results  were 
verified  by  Mr.  F.  R.  Low,  editor 
of  "Power"  (S.  page  160). 

In  Fig.  45  is  shown  a  small 
vertical  una-flow  engine  of  30  HP 
at  400  r.  p.  m.  for  marine  lighting 


service. 
Fig.  42.  A  better  design   is   shown  in 

Fig.  46,  illustrating  a  similar  engine 

of  30  HP  designed  by  the  author  for  an  English  firm.  The  cylinder  bore  is  220  mm, 
the  stroke  160  mm,  and  the  speed  400  r.  p.  m.  The  governor  eccentric  oscillates  a  roller 
lever  acting  on  a  triangular  cam  which  transmits  the  motion  to  the  valves.  The 
whole  cam  mechanism  is  enclosed  in  a  separate  housing  filled  with  oil.  The  cylinder 
ends  are  jacketed,  and  the  additional  clearance  pockets  are  formed  in  the  cylinder 
heads  and  arranged  to  be  heated  by  live  steam  when  operating  condensing.  The 
perfect  tightness  of-  the  single-beat  valves  employed,  the  small  clearance  space 
and 'clearance  surfaces,  the  generous  jacketing,  and  the  ample  exhaust  port  area 
all  combine  with  the  una-flow  action  to  insure  high  economy.  The  single-beat 
valves  provide  absolute  safety  against  damage  from  water  which  may  be  trapped 
in  the  cylinder. 

The  design  of  the  two  cylinder  vertical  single-acting  stationary  engine  shown 
in  Fig.  47  is  worthy  of  notice.    The  inlet  valves  are  single-beat  and  are  placed  in 


161 


the  center  of  the  cylinder  heads.  The  valve 
gear  consists  of  a  cam  mechanism  with 
reciprocating  slides  operated  through  a  bell 
crank  by  an  eccentric  and  shaft  governor 
located  at  the  free  end  of  the  crank  shaft. 
The  cranks  are  set  at  180°  in  order  to  obtain 
proper  balance  at  the  high  speed  at  which 
this  engine  is  t<5  run.  Single-beat  valves  are 
permissible  because  the  high  compression 
balances  the  pressure  against  which  they 
open.  The  valve  gear  parts  may,  however, 
be  easily  made  strong  enough  to  withstand 
the  load  if  the  valve  should  be  lifted  when 
there  is  no  compression  to  balance  the 


Fig.  43. 

pressure  upon  it.  The  cylinder  head  proper 
is  a  thin  dished  steel  plate,  and  the  cylin- 
der is  provided  with  a  forged  steel  liner. 
The  cylinder  head  is  jacketed  and  the 
upper  end  of  the  cylinder  is  also  heated  by 
live  steam  admitted  to  a  number  of  turned 
grooves.  Each  groove  communicates  with 
the  adjacent  ones  at  opposite  sides  so  that 
a  continuous  flow  may  take  place  through 
the  grooves.  The  engine  is  intended  for  use 
with  saturated  steam,  for  which  reason  the 
unjacketed  part  of  the  cylinder  next  to  the 
exhaust  belt  is  short.  The  forged  steel 
cylinder  liner  should  be  made  of  hard  ma- 
terial in  order  to  insure  satisfactory  ser- 
vice of  the  cast  iron  piston  rings.  The 
top  surface  of  the  piston  is  arched  to 

Slumpf,  The  una-flow  steam  engine. 


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163 

provide  sufficient  strength  with  a  light  section.  The  small  thickness  of  cylinder 
head  and  cylinder  liner  is  intended  to  bring  the  temperature  of  the  harmful  sur- 
faces as  close  as  possible  to  that  of  the  live  steam,  in  order  to  reduce  their  tempera- 
ture variation.  This  design  allows  the  additional  harmful  surface  to  be  reduced 


Fig.  45. 

to  as  small  an  amount  as  8%.  The  flow  of  steam  through  this  cylinder  takes  place 
in  such  a  perfect  manner  as  cannot  be  attained  by  any  other  cylinder  design.  The 
steam  enters  centrally  at  the  top,  spreads  out  in  all  directions  in  the  narrow  space 
between  cylinder  head  and  piston,  and  leaves  in  a  similarly  even  manner  at  the 

ll* 


164 

bottom  of  the  cylinder.    The  water  of  condensation  collects  on  the  piston  surface 
during  the  outstroke  and  is  completely  removed  by  the  rush  of  exhausting  steam, 


this  action  being  assisted  by  the  arched  form  of  the  piston  head.  The  una-flo\v 
principle,  together  with  the  very  favorable  flow  conditions,  the  thorough  draining 
of  the  cylinder  at  each  stroke,  the  ample  jacketing,  the  perfect  tightness  of  the 


165 

inlet  valve,  the  small  clearance  volume  and  surfaces  will  insure  a  very  low  steam 
consumption  with  this  type  of  engine. 

Proper  attention  must  be  paid  to  the  design  of  the  lower  part  of  the  piston 
which  forms  a  seal  against  vacuum. 


Fig.  47. 


Fig.  48  shows  an  interesting  cylinder  design  with  automatic  auxiliary  exhaust 
valves.  (See  also  chapters  on  withdrawal  of  sfeteam  and  on  locomobiles.)  This  engine 
is  a  una-flow  engine  in  a  restricted  sense  only,  since  at  light  loads  the  exhaust  steam 
leaves  through  the  valves  only,  while  for  longer  cut-offs  part  of  it  exhausts  also 
through  the  ports.  The  ports  in  the  cylinder  leading  to  the  auxiliary  exhaust 
valves  are  overrun  by  the  piston,  thus  determining  the  compression.  The  auxiliary 


166 


exhaust  valves  in  this  design  are  operated  automatically  by  the  working  steam 
of  the  cylinder;  they  may,  however,  be  opened  and  closed  by  a  separate  valve  mo- 
tion. On  the  stem  of  each  exhaust  valve  is  mounted  a  piston  working  in  a  cylinder 


Fig.  48. 


forming  part  of  the  valve  bonnet,  the  upper  side  of  which  is  connected  to  the 

corresponding  end  of  the  engine  cylinder.   When  the  main  piston  is  near  the  dead 

center,    the    high  pressure   in 

the  engine  cylinder   also   acts 

on  the  piston  attached  to  the 

auxiliary    exhaust    valve   and 

thus    holds   the  latter   closed 

until  the  main  exhaust  ports 

are  uncovered  and  the  pressure 

is  released.   The  spring  on  the 

upper  end  of  the  stem    then 


opens  the  valve  and  holds  it 


Fig.  49. 


167 


open  until  at  the  end  of  the  stroke  the  steam  pressure  rises  sufficiently  to  close  it. 
Exhaust  therefore  takes  place  through  the  valves  until  the  piston  overruns  the 
auxiliary  exhaust  ports.  The  clearance  space  in  this  design  is  not  larger  than 
that  of  an  ordinary  condensing  engine,  and  the  length  of  compression  can  be  chosen 

to  suit  any  required  conditions.    The  volume  loss 
will  be  considerably  reduced,   although  there  will 


Fig.  51. 


Fig.  50. 


Fig.  52. 


Fig.  53. 


be  a  small  increase  in  the  surface  loss.  In  order  to  keep  the  latter  down  to 
a  minimum,  the  auxiliary  exhaust  valves  should  be  placed  on  the  cylinder  barrel 
instead  of  in  the  heads,  and  arranged  so  that  in  the  dead  center  position  of  the 
piston  at  least  one,  or  preferably  two  rings  seal  the  auxiliary  exhaust  ports. 


168 


If  this  is  done,  the  compression  steam,  assisted  later  by  the  live  steam,  will  close 
the  valves  before  the  piston  uncovers  their  ports  on  the  expansion  stroke,  whereby 
direct  exhaust  of  live  steam  is  avoided. 

The  auxiliary  exhaust  has  no  value  for  condensing  engines,  but  some  value 
for  non-condensing  una-flow  engines  with  high  back  pressure  and  low  initial 
pressure,  where  it  will  result  in  a  reduction  in  steam  consumption,  especially 
in  the  case  of  jacketed  cylinders,  with  the  further  advantage  of  higher  mean 
effective  pressures,  and  therefore  smaller  cylinders,  driving  parts  and  flywheels. 

Figs.  49,  50,  51,  52,  53,  54,  55  refer  to  una-flow  non-condensing  engines 
built  by  the  Skinner  Engine  Company,  of  Erie,  Pa«,  U.S.A.  The  diagram  (Fig.  49) 


<t  or  VALVE  STEM 
WHEN  RUNNING  WTTH^Ij 
LIGHT  LOAD 


of  VALVE  STEM 

WHEN   RUNNING   WITH 

Fuu.  LOAD 


AM  ROCKER 
ANDCA.M  WHEN 

RUNNING 

OR  FULL  LOAD 


Fig.  54.     Sectional  View  of  Expansion- Compensating  Gear. 

shows  delayed  compression  by  the  use  of  auxiliary  exhaust  valves  (Fig.  50)  arranged 
at  an  intermediate  point  of  the  stroke.  The  clearance  space  may  be  reduced  as 
much  as  can  be  realized  by  the  best  practical  design,  and  the  ports  leading  to  the 
valves  may  be  arranged  so  as  to  give  a  compression  complying  with  the  rules  deve- 
loped on  page  41,  42,  43.  Single-beat  valves  are  employed,  since  the  pressure 
upon  them  is  relieved  before  they  open,  by  the  piston  uncovering  the  main  exhaust 
ports.  The  auxiliary  exhaust  valves  are  arranged  on  the  lower  side  of  the  cylinder 
for  drainage.  The  inlet  valves  are  placed  on  the  upper  side  of  the  cylinder,  and 
are  constructed  as  double-beat  valves  (Fig.  51)  in  accordance  with  the  principles 
utilized  in  the  Sulzer  valve  shown  in  Fig.  10,  p.  88.  The  upper  valve  face  is  formed 
on  a  disc  separate  from  the  main  valve  body,  and  the  two  parts  are  pressed  together 
by  a  spring,  thus  allowing  a  limited  amount  of  movement  between  them  so  that 
the  faces  can  adapt  themselves  to  a  change  in  distance  between  the  fixed  seats 
due  to  expansion.  Snap  rings  are  fitted  between  the  two  parts  of  the  valve  so  as 
to  seal  the  joint.  An  eccentric  controlled  by  a  flywheel  governor  operates  a  rocking 


169 


Fig.  55. 


Fig.  56. 


170 


bo 
fe 


171 


shaft  placed  between  or  alongside  the  inlet  valves,  the  latter  being  actuated  by 
cams  and  levers.  The  axial  arrangement  of  the  rocking  shaft  entirely  eliminates 
the  effect  of  the  expansion  of  the  cylinder  on  the  steam  distribution.  The  cam 

mechanism  of  both  inlet  and  exhaust  valves  is  enclosed  in  an 

oil  bath. 


Fig.  60. 

In  Fig.  50  is  shown  a  device  to  adapt  the  engine  auto- 
matically to   both   condensing  and  non-condensing  operation. 
The  shaft  A   carries   an  idler  lever   B  which  is  actuated   By 
the   rolling  lever   C.     The  latter  is   operated  by  the   engine  valve  gear  through 
the  shaft  D  on  the  outside  of  the  cam  housing.     The  pocket  E  is  connected  to 
the  central  exhaust  belt  by  means  of  a  small  pipe.     A  spring  in  this  pocket  bears 
on  the  idler  shaft  so  as  to  keep  the  idler  lever  in  register  with  the  valve  stem  and 
the  rolling  lever  C  below  it.    When  the  vacuum  increases  sufficiently,  the  shaft 


172 

is  drawn  into  the  pocket  against  the  spring,  so  that  the  idler  and  rolling  levers  no 
longer  register  and  the  valve  remains  closed.  When  the  vacuum  fails,  the  spring 
will  cause  the  levers  to  fall  in  line,  so  that  the  valve  becomes  operative. 


Fig.  61. 


The  Nordberg  Mfg.  Co.,  of  Milwaukee,  Wis.,  was  the  first  American  concern 
to  take  up  the  manufacture  of  una-flow  engines  with  the  constructional  features 
proposed  by  the  author,  combined  with  a  design  according  to  their  own  practice 
(Fig.  56).  The  shaft  gavernor  on  the  lay  shaft  controls  the  valve  mechanism  in 
the  usual  way.  The  valves  are  positively  opened  and  closed  by  cams,  in  accordance 
with  the  design  originated  by  Prof.  Doerfel,  of  Prague  (Fig.  57  and  Fig.  58).  '  No 
springs  or  dashpots  are  required,  except  a  short  spring  inserted  in  the  connection 
between  valve  and  valve  stem,  to  insure  proper  closure  of  the  valve. 


Fig.  62. 

A  synchronizing'  device  is  regularly  furnished  with  all  engines  driving  alter- 
nating current  generators.  A  hand  wheel  is  arranged  at  the  end  of  the  lay  shaft, 
by  means  of  which  the  tension  of  the  governor  springs  may  be  varied  while 
the  engine  is  running  to  change  the  speed  and  to  bring  the  generator  into 
synchronism. 

It  should  be  noted  that  a  tail  rod  and  slipper  are  used  (Fig.  59).  In  larger 
engines  the  piston  rod  is  made  hollow. 


173 


The  crosshead  (Fig.  60)  carries  a  pin  with  tapered  ends  pulled  into  place  by 
a  large  nut.  The  crosshead  shoes  are  centered  between  the  projecting  flanges  of 
the  crosshead  body,  are  faced  with  suitable  babbitt  metal  and  provided  with 
a  wedge  and  screw  adjustment. 


<0 

bb 


to 

be 


The  crank  end  of  the  connecting  rod  is  of  especial  interest  (Fig.  61).  The 
cylindrical  stationary  box  is  fitted  into  the  bored  eye  of  the  rod.  The  adjustable 
box  is  slightly  narrower  than  the  diameter  of  the  pin  and  fits  into  a  recess  in 
the  rod.  One  side  of  the  stationary  box  is  cut  away,  forming  a  slot  through  which 
the  adjustable  box  projects,  its  concave  surface  bearing  against  the  pin.  The 
stationary  box  is  thus  prevented  from  pinching  the  pin  and  from  moving  with  the 


174 

latter.  The  adjusting  wedge  has  the  form  of  a  cylindrical  block  of  steel,  one  side  of 
which  is  cut  away  to  form  a  plane  surface  inclined  to  its  axis  and  bearing  against 
the  corresponding  inclined  surface  on  the  adjustable  box.  The  wedge  fits  into 


Fig.  65. 


Fig.  66. 

a  reamed  hole  in  the  rod  and  is  adjusted  ba  cap  screws  locking  each  other.  The 
small  end  of  the  rod  is  fitted  with  boxes  of  the  same  design  (Fig.  62).  This  design 
combines  simplicity  of  manufacture  with  thorough  reliability  in  operation. 


175 


to 
be 


Fig.  69. 


176 


The  frame  (Figs.  63,  64)  is  cast  in  one  piece  with  smooth  surfaces  and  straight 
outlines.  The  center  is  kept  as  near  to  the  foundation  as  possible  and  a  great  width 
is  provided  at  the  front  of  the  guides,  where  the  bending  moment  is  a  maximum. 
The  frames  are  cast  base  uppermost,  thus  insuring  good  clean  metal  in  the  guides 


o 
r> 

be 


be 


and  the  line  of  stress.    Provision  is  also  made  in  the  design  for  the  collection  of 
the  oil  and  its  return  to  the  lubricating  system. 

The  moving  parts  of  the  engine  are  oiled  by  a  gravity  overhead  lubricating 
system.  The  cylinder  is  lubricated  by  a  mechanical  force  feed  lubricator  distri- 
buting oil  positively  to  the  proper  points. 


177 

The  main  bearings  (Figs.  65  and  66)  are  of  the  quarter-box  type,  lined  with 
babbitt  metal.  The  cap  forms  a  strong  tie  and  is  relieved  in  the  middle  so  that  it 
exerts  pressure  directly  over  the  quarter  boxes.  In  the  construction  shown  in  Fig.  65, 
the  adjustment  is  effected  by  means  of  heavy  set  screws,  which  are  provided  with 
steel  contact  blocks  to  prevent  them  from  wearing  into  the  quarter  boxes,  while 
in  Fig.  66  the  same  effect  is  obtained  by  vertical  wedges  operated  by  screws  passing 
through  the  cap. 

The  Nordberg  Mfg.  Co.  also  builds  una-flow  engines  with  auxiliary  exhaust 
valves  which  may  be  put  in  or  out  of  operation  thus  making  the  engines  suitable 
for  both  condensing  and  non-condensing  service. 


Fig.  72. 


Owing  to  careful  use  of  those  principles  which  have  proved  successful  in  Europe, 
especially  those  developed  by  the  author,  the  una-flow  engines  built  by  the  Nord- 
berg Mfg.  Co.  may  be  considered  among  the  best  American  machines  of  this  type. 
Their  first  engine  was  fitted  with  Corliss  valves  as  described  in  chapter  2 — 2,  page  184. 
Still  further  improvement  in  these  engines  might  be  made  by  the  application  of 
high  lift  single  beat  valves  without  cages,  thereby  reducing  the  clearance,  the 
harmful  surfaces,  and  leakage,  so  that  the  water  rate  might  be  still  further  im- 
proved. 

The  dual  clearance  una-flow  engine  built  by  the  Harrisburg  Foundry  &  Ma- 
chine Works,  of  Harrisburg,  Pa.  (Fig.  67)  is  of  especial  interest.  As  the  name 

Stumpf,  The  una-flow  steam  engine.  12 


178 

implies,  this  engine  has  two  clearances,  the  first,  or  cylinder  clearance,  between 
the  piston  and  the  valve,  and  the  second  (EGE,  Fig.  68)  connecting  the  outer  ends 
of  the  valve.  In  the  earlier  design  this  connection  was  by  an  external  pipe  (Fig.  68) 
but  a  hollow  valve  is  now  used  (Fig.  71).  The  latter  is  an  ordinary  piston  valve 
with  snap-rings,  moving  in  a  ribbed  bushing  and  having  inside  admission,  as  in 
locomotive  practice  with  superheated  steam.  As  indicated  in  the  figure,  the  cy- 
linder clearance  is  kept  as  small  as  possible.  The  residual  steam  is  first  compressed 
into  both  clearance  spaces  together,  which  are  at  that  time  in  connection  through 
the  piston  valve.  Towards  the  end  of  the  stroke  the  auxiliary  clearance  is  shut 
off,  so  that  the  steam  is  compressed  into  the  cylinder  clearance  alone.  The  valve 
then  automatically  connects  the  auxiliary  clearance  with  the  opposite  end  of  the 
cylinder  so  that  the  steam  compressed  into  the  clearance  now  mixes  with,  and 
expands  with  the  working  steam  on  the  return  storke.  This  arrangement  is  of 
course  only  justified  for  non-condensing  service,  so  as  to  avoid  a  large  clearance 
or  auxiliary  exhaust  valves.  The  una-flow  principle  is  fully  adhered  to,  since  the 
exhaust  steam  only  leaves  through  the  central  exhaust  ports.  The  series  arrange- 
ment of  live  steam,  inlet  valve,  piston  and  exhaust  is  also  retained,  so  that 
any  leakage  of  steam  past  the  piston  valve  cannot  pass  directly  to  the  exhaust. 
The  effect  of  this  dual  clearance  principle  is  to  raise  the  expansion  line  and 
lower  the  compression  line  (Fig.  69).  Consequently  the  mean  effective  pressure, 
output  and  uniformity  of  speed  are  somewhat  increased  and  the  necessary 
flywheel  weight  is  decreased. 

Conditions  in  an  engine  of  this  type  for  condensing  service  are  somewhat 
different.  The  piston  heads  are  flat,  but  in  spite  of  this  there  remains  a  clearance  of 
from  5°/0  to  7%.  In  this  case  also  the  auxiliary  clearance  connects  the  valve  head 
pockets.  In  addition,  a  further  clearance  pocket  is  arranged  to  be  connected  to  the 
cylinder  by  a  spring-loaded  valve,  which  may  be  opened  or  closed  by  hand,  or 
operated  automatically,  thereby  adapting  the  condensing  cylinder  to  non-conden- 
sing service  (Fig.  72).  Increased  compression  will  cause  the  valve  to  open  against 
the  spring,  thus  relieving  the  pressure  by  admitting  steam  into  the  pocket. 

The  heads  and  ends  of  the  cylinder  barrels  are  jacketed  for  both  condensing 
and  non-condensing  service.  The  piston  valve  is  actuated  in  both  cases  by  an 
eccentric  on  the  crankshaft  controlled  by  a  shaft  governor. 

The  Filer  &  Stowell  Co.,  of  Milwaukee,  Wis.,  successfully  built  a  una-flow 
engine  with  drop  piston  valves  and  a  valve  gear  resembling  a  Corliss  gear,  the 
valves  being  located  at  the  side  of  the  cylinder  barrel.  In  the  later  type,  the  valves 
were  placed  in  the  heads  on  the  cylinder  center  line,  thus  decreasing  the  clearance 
and  making  a  better  jacket  arrangement  possible.  For  higher  engine  speeds,  poppet 
valves  were  adopted  later,  operated  from  eccentrics  on  a  lay  shaft,  with  a  positive 
opening  and  closing  motion,  the  eccentrics  being  controlled  by  a  lay  shaft  governor 
placed  between  them.  The  result  is  an  engine  similar  in  many  respects  to  the  Nord- 
berg  construction,  the  chief  points  of  difference  being  the  valve  gear  and  bonnet 
design.  The  Filer  &  Stowell  Go.  evidently  consider  thorough  jacketing  of  much 
importance,  and  therefore  their  claim  of  a  consumption  of  137/s  Iks.  °f  saturated 
steam  per  I  HP  per  hour  for  a  una-flow  engine  having  a  16  X  30"  cylinder,  125 
Ibs./sq.  in  initial  pressure,  25"  vacuum  and  150  r.  p.  m.  may  well  be  credited.  For 


CO 

r^ 

tic 


12' 


181 

these  engines  also  the  next  logical  step  would  be  to  adopt  the  single-beat  valves 
actuated  from  a  double-speed  lay  shaft. 

Fig.  73  illustrates  a  Filer  &  Stowell  20  X  22"  una-flow  engine  driving  a 
200  kW  direct  current  generator.  This  engine  is  arranged  for  condensing  and  non- 
condensing  service,  auxiliary  valves  being  employed  in  preference  to  clearance 
spaces  because  of  more  or  less  extended  periods  of  operation  with  as  high  a  back 
pressure  as  5  Ibs/sq  in.  The  clearance  space  formed  in  each  head  by  the  auxiliary 
exhaust  valve  pocket  may  be  closed  off  by  a  hand-operated  clearance  valve,  so 
that  no  clearance  is  added  when  running  condensing.  If,  for  some  reason,  the 
vacuum  should  drop  to  20"  or  22",  the  hand-operated  clearance  may  be  opened, 
thus  adding  the  clearance  formed  by  the  exhaust  valve  pocket,  the  valve  itself 
remaining  closed.  If  the  back  pressure  is  further  increased,  the  mechanism  actu- 
ating the  exhaust  valves  may  be  put  in  action  and  the  length  of  compression  varied 
to  suit  the  back  pressure  while  the  engine  is  in  operation. 

Fig.  74  shows  a  group  of  five  18  X  42"  una-flow  cylinders  which  are  part 
of  an  order  of  six  18  X  42"  and  six  20  X  48"  cylinders.  Four  of  the  18  X  42" 
cylinders  are  to  take  the  place  of  those  of  two  18"  and  36  X  42"  cross-compound 
corliss  engines  driving  generators,  and  the  two  others  are  to  replace  those  of  two 
ammonia  boosters.  The  six  20  X  48"  cylinders  are  to  supplant  cross-compound 
cylinders  driving  ice  machines.  All  these  engines  are  installed  at  Swift  &  Company's 
plant  at  La  Plata,  Argentine  Republic,  and  are  to  operate  with  175  Ibs/sq  in.  steam 
pressure  and  150°  superheat.  All  these  cylinders  have  additional  clearance  spaces 
and  clearance  valves  for  non-condensing  service.  The  trend  of  development  is 
thus  the  same  as  in  Europe,  where  many  old  compound  cylinders  are  being  replaced 
by  single-stage  una-flow  cylinders. 


182 


2.  The  Corliss  Una-Flow  Engine. 

Figs.  3  and  1  show  a  side  elevation,  a  vertical  section  and  an  indicator  card 
of  a  condensing  Corliss  una-flow  engine.  The  eccentric  on  the  crank  shaft  directly 
operates  the  two  valves  placed  in  the  cylinder  heads.  The  release  is  effected  in 
the  ordinary  way  by  a  cam  under  control  of  the  governor.  The  valves  are  closed 

by  oil-vacuum  dash  pots  (Fig.  2)  which 
have  absolutely  no  rebound  after  the 
valve  has  closed.  The  oscillation  which 
is  so  common  with  air  dash  pots  is 
entirely  avoided,  since  air  which  is  a  com- 
pressible medium  is  replaced  by  oil  which 
is  incompressible.  For  this  reason  the  lap 
of  the  valve  can  be  made  extremely  small 


Q 


Fig.  1. 


Fig.  2. 


and  the  latch  only  takes  hold  and  begins  to  move  the  valve  at  a  time  when  the 
latter  is  already  partly  balanced  by  the  compression  (Fig.  1).  Further  experience 
is  required  to  find  the  smallest  allowable  lap  of  the  valve  in  combination  with  the 
proper  construction  of  a  reliable  oil  vacuum  dash  pot  which  exactly  locates  the 


184 

valve  in  its  end  position,  in  order  to  reduce  as  far  as  possible  its  movement  when 
unbalanced.  In  this  manner  it  should  be  possible  to  obtain  reliable  operation  with 
high  pressures  and  superheat.  Provision  must  be  made  in  the  valve  gear  to  close 
the  valve  positively  in  case  the  force  of  the  oil  dash  pot  should  be  insufficient  for 
proper  closure,  owing  to  very  small  cut-offs  or  other  reasons. 

A  negative  angle  of  advance  of  45°  gives  an  ample  range  of  cut-off  and  also 
insures  release  under  all  conditions,  if  the  governor  connections  are  properly 
adjusted.  The  omission  of  the  wrist  plate  and  the  attainment  of  extremely  small 
clearance  volumes  and  surfaces  are  further  valuable  features  of  this  design. 

Fig.  4  shows  a  Corliss  una-flow  cylinder  as  constructed  by  the  Nordberg 
Manufacturing  Company,  of  Milwaukee,  Wis.  In  accordance  with  the  above  prin- 
ciples, the  inlet  valves  have  very  small  lap  and  are  multiported  in  addition,  so 


Fig.  4. 

that  four  edges  open  simultaneously.  In  this  way  the  rotative  movement  of  the 
valve  is  reduced  and  high  speeds  are  rendered  possible.  The  additional  clearance 
volume  and  clearance  valves  for  non-condensing  operation  are  placed  in  the  lower 
part  of  the  cylinder  heads. 

In  Fig.  5  is  shown  a  non-condensing  Corliss  engine  with  the  inlet  valves  arranged 
in  the  heads,  the  auxiliary  exhaust  valves  at  the  ends  of  the  cylinder  barrel,  and 
the  una-flow  exhaust  ports  in  the  center. 

The  eccentric  on  the  crankshaft  has  a  negative  angle  of  advance  of  15°  and 
drives  the  inlet  valves  directly  through  an  ordinary  Corliss  mechanism,  while  the 
motion  of  the  exhaust  valves  is  derived  from  that  of  the  inlet  valve  levers.  The 
inlet  valves  have  the  ordinary  releasing  gear.  Admission  and  cut-off  are  deter- 
mined by  the  steam  valves,  the  beginning  of  exhaust  is  fixed  by  the  piston  unco- 
vering the  central  exhaust  ports  and  compression  is  delayed  until  the  auxiliary 
exhaust  ports  are  overrun  on  the  return  stroke.  The  valve  gear  proper  therefore 


185 


Fig.  5. 


186 

has  no  influence  whatever  upon  the  exhaust  phases.  This  renders  possible  the 
use  of  a  negative  angle  of  advance  and  therewith  a  range  of  cut-off  of  from  0  to 
more, than  60%.  The  length  of  compression  may  be  reduced  to  about  8%  of  the 
stroke  in  connection  with  the-  extremely  small  clearance  volume  realized  with  this 
design.  The  auxiliary  exhaust  valves  are  protected  by  the  piston  rings  against 
high  pressures  and  temperatures  during  a  considerable  part  of  the  admission 
period.  They  are  subject  to  pressure  only  after  the  piston  has  uncovered  the 
auxiliary  ports,  and  do  not  open  until  the  outer  dead  center  is  reached,  when  the 
main  exhaust  ports  are  wide  open.  Closure  of  the  valves  takes  place  after  the 
piston  has  again  covered  the  auxiliary  ports.  The  actual  closure  of  the  auxiliary 
exhaust  is  therefore  so  rapid  that  the  indicator  card  shows  sharp  corners  at  this 
point.  Since  the  inlet  valves  also  close  quickly,  and  as  the  clearance  volume  and 
surfaces  are  small,  a  low  steam  consumption  may  be  expected  with  this  type  of 
engine. 


187 


3.  The  Una-Flow  Engine  arranged  for  Bleeding. 

Steam  may  be  withdrawn  or  bled  from  a  una-flow  cylinder  by  means  of  check 
valves  placed  at  suitable  points,  as  shown  in  Fig.  1.  For  instance,  by  providing 
ports  at  a  distance  of  say  10%  of  the  piston  stroke  from  the  opening  edge  of  the 
main  exhaust  ports,  the  steam  withdrawn  through  the  former  may  be  used  in  heating 
systems,  to  drive  exhaust  steam  turbines  or  engines,  or  for  other  purposes.  It  is 
possible,  for  example,  to  withdraw  steam  at  a  pressure  of  0,5  to  1,0  at.  abs.  from 
the  cylinder  of  a  condensing  engine  and  to  use  it  in  an  exhaust  steam  turbine 
driving  a  rotary  air  pump.  The  bleeder  valves  could  also  be  placed  closer  to  the 


-  -1- 

\  I 


Fig.  1. 


188 

cylinder  ends  in  order  to  withdraw  steam  at  a  higher  pressure  for  heating  purposes 
and  the  like.  A  plurality  of  such  heating  systems  may  thus  be  arranged  in  series, 
as  shown  in  Fig.  1,  for  heating  the  feed  water  to  a  high  temperature.  In  this  case, 
however,  a  more  or  less  noticeable  loss  of  diagram  area  must  be  expected. 


Fig.  2. 


An  example  of  such  withdrawal  of  steam  is  shown  in  Fig.  2,  which  illustrates 
a  una-flow  locomotive  cylinder  fitted  with  automatic  bleeder  valves.  In  locomo- 
tives this  method  of  withdrawing  steam  may  very  well  be  considered  for  train 
heating  or  feed  water  heating.  The  diagrams  of  Figs.  3  and  4  give  the  quanti- 


°g\100  SO  SO  VO  2O 


iff,  25% 


Fig.  3. 


ties  which  may  be  withdrawn  at  different  locomotive  speeds  with  valves  placed 
at  35  and  50%  of  the  stroke  before  the  dead  center.  The  quantities  of  steam  with- 
drawn are  the  larger,  the  greater  the  area  of  the  valves,  the  greater  their  distance 


189 


from  the  exhaust  end,  the  lower  the  speed  and  the  lower  the  required  pressure. 
The  following  table  gives  the  quantities  of  steam  withdrawn  for  different  engine 
speeds  and  pressures  of  the  heating  steam,  the  bleeder  valves  being  situated  either 
35  or  50%  of  the  stroke  from  the  outer  dead  center.  The  figures  denote  quantities 
of  steam  withdrawn  in  percent  of  the  total  working  steam  in  the  cylinder. 

Steam  quantities  withdrawn: 


Pressure 
in  receiver 
kgs.  per  cm1 

Cut-off  at  quarter  of  the  stroke 

45  km  per  hour 

60  km  per  hour 

75  km  per  hour 

90  km  per  hour 

Withdrawal  valve  35%  before  the  end  of  the  stroke: 


1.25 

59.0% 

55.0% 

52.0% 

43.0  % 

1.5 

51.0 

48.5 

46.0 

41.0 

2.0 

37.5 

35.2 

32.5 

30.0 

2.5 

25.0 

22.5 

20.0 

17.0 

Withdrawal  valve  midway  in  the  stroke: 


2.5 

41.8 

39.4 

37.0 

33.6 

3.0 

29.4 

27.0 

25.1 

22.8 

3.5 

20.1 

18.7 

17.0 

15.3 

4.0 

11.3 

10.3 

9.2 

7.9 

clearance  s/iace 
fff.zs 


Fig.  4. 


190 

The  amount  of  steam  which  can  be  withdrawn  from  the  cylinder  of  a  una- 
flow  condensing  engine  will  of  course  be  considerably  smaller  than  in  a  non-con- 
densing locomotive. 

This  very  important  problem  of  bleeding  steam  may  also  be  solved  by  means 
of  a  compound  una-flow  engine  with  a  tandem  arrangement  as  shown  in  Fig.  5. 
The  piston  of  the  high  pressure  una-flow  cylinder  is  fitted  with  a  piston  valve  driven 
from  the  main  connecting  rod,  which  provides  the  necessary  short  compression. 
The  same  effect  may  be  obtained  by  the  use  of  automatic  auxiliary  exhaust  valves 
operated  by  the  cylinder  steam  and  placed  near  the  ends  of  the  cylinder,  their 
ports  being  controlled  by  the  piston  (Fig.  6).  The  low  pressure  cylinder  is  of  stan- 
dard una-flow  construction,  although  its  clearance  volume  must  be  increased  to 
about  7  to  10%  according  to  the  receiver  pressure.  The  cut-off  of  the  low  pressure 
cylinder  may  be  controlled  by  a  pressure  regulator  under  the  influence  of  the  re- 
ceiver pressure  (Fig.  7).  This  regulator  consists  of  a  cast  iron  housing  partly  filled 


Fig.  5. 

with  mercury  carrying  an  iron  float,  the  upper  end  of  which  is  connected  in  a  sui- 
table way  to  the  inlet  gear  of  the  low  pressure  cylinder.  The  position  of  this  float 
changes  with  the  height  of  the  mercury  column  displaced  by  the  receiver  pressure. 
The  connection  is  made  in  such  a  way  that  the  regulator  shortens  the  cut-off  of 
the  low  pressure  cylinder  for  a  decreasing  receiver  pressure.  In  the  arrangement 
shown  in  Fig.  7  the  pressure  regulator  acts  upon  an  intermediate  pin  in  the  drive 
from  the  eccentric  to  the  inlet  valves,  in  such  a  way  as  to  displace  the  rods  from 
a  straight  line  to  a  position  on  either  side  of  it,  thus  producing  the  desired  change 
of  cut-off,  within  certain  limits,  with  a  permissible  change  in  lead.  The  piston 
of  the  auxiliary  exhaust  valve  (Fig.  6)  is  under  the  influence  of  the  cylinder  pres- 
sure. The  valve  will  therefore  be  closed  whenever  the  pressure  inside  the  cylinder 
is  high.  It  is  opened  by  a  spring  in  the  valve  bonnet  when  the  main  piston  uncovers 
the  exhaust  ports  and  also  when  the  expansion  reaches  the  back  pressure  before 
this  occurs.  The  losses  accompanying  a  loop  in  the  exhaust  line  of  the  indicator 
card  for  small  cut-offs  are  therefore  avoided.  It  is  also  possible  to  use  large  high 
pressure  cylinders  even  for  long  cut-offs  and  yet  avoid  a  large  pressure  drop  at 
the  end  of  expansion. 

The  opening  and  closing  of  these  automatic  valves  is  rendered  noiseless  by  an 
adjustable  double-acting  oil  dash  pot  (Fig.  48,  ch.  II,  1,  p.  166).  The  lower  resilient 
seat  insures  tightness  of  the  valve.  The  design  permits  of  ample  valve  lift  and 


valve  areas,  quiet  operation  and  small  clearance,  as  well  as  very  favorable  sealing 
conditions  during  the  first  and  last  part  of  the  stroke,  when  the  piston  rings  come 
between  the  inlet  and  auxiliary  exhaust  valves.  Further  valuable  features  of  this 
design  are  found  in  the  arrangement  of  all  the  valves  and  their  gear  on  top  of  the 
cylinder,  and  the  unhampered  disposal  of  all  the  piping  together  with  the  condenser 
underneath  the  same,  so  that  every  pipe  flange  is  easily  accessible. 


Fig.  6. 

The  exhaust  valves  of  the  high  pressure  cylinder  may  also  be  arranged  in  the 
more  usual  way  with  separate  valve  gear,  possibly  with  omission  of  the  central 
exhaust. 

In  Fig.  8  is  shown  an  arrangement  which  makes  it  possible  to  withdraw  steam 
during  both  the  expansion  and  compression  strokes.  The  una-flow  cylinder  of 
standard  design  is  fitted  at  each  end  with  an  automatic  auxiliary  exhaust  valve 
controlled  by  the  cylinder  steam  as  previously  described.  The  upper  end  of  the 
stem  of  the  balanced  valve  carries  a  piston  working  in  a  cylinder  the  upper  side 
of  which  is  connected  by  a  pipe  with  the  corresponding  end  of  the  engine  cylinder. 
This  pipe  connection  is  fitted  with  a  reducing  valve  which  opens  towards  the  valve 
cylinder  when  a  certain  pressure  is  reached.  The  same  end  of  the  valve  cylinder 
also  has  a  second  pipe  connection  leading  to  a  pilot  valve  communicating  with  the 


192 


condenser,  which  is  operated  by  a  mechanism  driven  by  the  layshaft.  When  this 
pilot  valve  is  lifted,  pressure  release  occurs  above  the  piston  of  the  auxiliary  exhaust 
valve,  and  the  latter  is  opened  by  its  spring,  thus  admitting  steam  from  the  engine 
cylinder  into  the  heating  connection.  When  the  pressure  inside  the  cylinder  falls 
to  or  below  that  carried  in  the  heating  system,  a  return  flow  of  steam  is  pre- 
vented by  the  closure  of  a  number  of  automatic  metal  strip  flap  valves  disposed 
around  the  auxiliary  exhaust  valve.  The  latter,  however,  remains  open  until  after 
the  auxiliary  exhaust  ports  are  covered  by  the  main  piston  on  its  return,  when 
the  rising  compression  pressure  acts  upon  the  valve  piston  through  the  pipe  con- 
nection previously  mentioned,  and  thus  closes  the  valves.  The  opening  or  timing 
of  the  small  pilot  valve,  which  is  operated  by  the  layshaft,  is  controlled  by  the 
pressure  in  the  heating  system  by  means  of  an  apparatus  containing  an  iron  float 
carried  on  mercury.  The  lower  mercury  level  is  exposed  to  the  pressure  in  the 

heating  system,  and  a  fall  of  pres- 
sure in  the  latter  will  therefore  lower 
the  mercury  column  and  change  the 
position  of  the  float,  thus  causing 
the  auxiliary  valve  to  open  earlier 
and  allowing  more  steam  to  be 
withdrawn  from  the  engine  cylinder. 
Conversely,  if  the  pressure  in  the 
heating  system  rises,  the  mercury 
column  will  also  rise  and  the  float 
in  its  new  position  will  open  the 
pilot  and  exhaust  valves  later  so 
that  less  steam  will  be  withdrawn. 
The  middle  position  of  the  float 
corresponds  to  a  horizontal  position 
of  the  shor^  link  connecting  the 
upper  end  of  the  float  rod  with  a 
small  crank,  and  the  motion  of  the 
latter  will  thus  be  the  same  whether 
the  float  rises  or  falls.  The  crank 
will  therefore  be  in  the  extreme  left 


Fig.  7. 


position  for  the  middle  position  of  the  float,  and  in  the  extreme  right  position  for 
either  the  highest  or  lowest  position  of  the  float.  This  crank  rocks  an  eccentric 
pivot  upon  which  is  mounted  the  double-armed  lever  operating  the  small  pilot 
valve.  These  levers  are  moved  by  eccentrics  keyed  to  the  lay-shaft. 

The  operation  of  the  whole  mechanism  will  be  clear  from  a  study  of  the  series 
of  indicator  cards  reproduced  in  Fig.  8.  Starting  with  the  highest  float  position, 
the  point  of  opening  of  the  auxiliary  exhaust  valve  moves  more  and  more  towards 
the  left  while  the  float  falls.  This  continues  until  the  float  reaches  its  middle 
position,  for  which  the  cut-off  in  the  engine  cylinder  determined  by  the  load  is 
so  short  that  hardly  any  steam  can  be  withdrawn  during  expansion.  (See  cards 
No.  2  and  1.)  The  regulating  mechanism  follows  the  change  of  cut-off.  If  much 
heating  steam  is  required  when  the  engine  is  operating  with  small  loads  the  float 


193 


, 


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be 


Stamp/,  The  una-flow  steam  engine. 


13 


194 

will  fall  still  further,  the  small  crank  again  moves  towards  the  right;  and,  since 
no  more  steam  is  available  during  expansion,  an  arrangement  comes  into  play 
by  which  steam  is  withdrawn  during  compression.  This  is  done  by  raising  the 
back  pressure  and  therewith  the  compression  line,  by  admitting  air  into  the  con- 
denser through  a  snifting  valve  actuated  by  a  connecting  link  from  the  upper 
end  of  the  float.  The  compression  steam  then  soon  attains  the  heating  pressure 
and  escapes  through  the  still  open  auxiliary  exhaust  valve,  and  past  its  check 
valves  (see  diagram  8). 

If  much  steam  is  required  while  the  engine  is  running  with  heavy  loads,  then 
withdrawal  occurs  during  both  expansion  and  compression,  as  is  shown  in  dia- 
grams 4,  5  and  6.  The  cut-off  in  this  case  is  late  enough  so  that  even  for  the 
lowest  float  positions  steam  will  be  withdrawn  during  expansion. 

The  diagrams  in  a  general  way  show  that  the  action  of  this  mechanism  tends 
to  produce  the  smallest  possible  loss  due  to  incomplete  expansion.  Instead  of 
using  a  crank  mechanism  for  operating  the  small  pilot  valves,  a  cam  may  be  fitted 
on  the  upper  end  of  the  float  rod  which  acts  upon  a  spring-loaded  roller  and  thus 
adjusts  the  position  of  the  fulcrum  of  the  double-armed  lever.  This  combination 
has  the  advantage  that  by  properly  designing  the  cam  profile  any  desired  depen- 
dence between  float  travel  and  exhaust  valve  timing  may  be  realized. 

In  order  to  obtain  sufficient  valve  area  for  the  withdrawal  of  steam  during 
expansion,  each  end  of  the  cylinder  is  provided  with  two  of  the  automatic  exhaust 
or  bleeder  valves,  each  surrounded  by  a  nest  of  check  valves.  The  end  of  tlie 
lay-shaft  carries  two  small  eccentrics  which  operate  the  pilot  valves  through  the 
above  mentioned  double-armed  levers  controlled  by  a  common  float.  The  governor 
on  the  lay-shaft  controls  the  engine  output  entirely  independently  of  the  heating 
requirements. 

So  far  the  problem  of  withdrawing  steam  for  heating  purposes  has  been  solved 
mostly  by  installing  a  tandem  engine  with  a  counterflow  high  pressure  and  una- 
flow  low  pressure  cylinder,  the  heating  steam  being  taken  from  the  receiver  as 
was  described  above. 


195 


4.  The  Una-Flow  Rolling  Mill  Engine. 

Much  credit  for  their  work  in  this  field  is  due  to  the  firm  of  Ehrhardt  &  Sehmer, 
who  originated  the  design  shown  in  Fig.  1.  The  latter  is  noteworthy  for  the  use 
of  valve  gear  of  the  Zvonicek  type,  which  has  the  advantage  of  ample  valve  opening 
at  short  cut-offs  as  well  as  a  large  range  of  admission  without  excessive  valve 
lifts  at  late  cut-offs.  The  governor  adjusts  the  cut-off  by  moving  the  eccentric 
strap  which  carries  a  cam  profile  at  its  upper  side  for  the  eccentric  rod  roller  to 
work  upon.  For  rolling  mill  engines  a  maximum  cut-off  of  50  to  60%  is  absolutely 
essential.  This  is  easily  obtainable  with  the  Zvonicek  gear,  the  only  disadvantage 
of  which  is  its  complication,  although  this  has  never  proved  to  be  a  source 
of  complaint.  The  single  stage  una-flow  engine  gives  considerably  more  power 
than  the  tandem  compound  and  it  will  pull  through  where  the  latter  would 
stall.  This  feature  is  highly  appreciated  by  rolling  mill  engineers  on  account 
of  the  varying  resistance  of  the  roljs,  and  is  the  main  reason  for  the  rapid  intro- 
duction of  the  una-flow  engine  in  rolling  mill  practice.  This  preference  has  even 
led  to  the  replacement  of  several  old  tandem  cylinders  by  cylinders  of  the  una- 
flow  type. 

Fig.  2  shows  a  flywheel  rolling  mill  engine  of  medium  size  built  by  Ehrhardt 
&  Sehmer.  The  engine  has  Stumpf  valve  gear,  and  the  condenser  air  pump  is  driven 
by  the  tail  rod. 

Figs.  3  and  4  show  the  cylinder  oLa  larger  una-flow  rolling  mill  engine  with 
Zvonicek  valve  gear  also  built  by  Ehrnardt  &  Sehmer. 

Fig.  5  shows  an  ordinary  tandem  reversing  rolling  mill  engine  built  by  Ehr- 
hardt &  Sehmer. 

The  three  pairs  of  cylinders  act  on  three  cranks  set  at  120°,  which  arrange- 
ment requires  less  maximum  admission  than  cranks  at  90°.  This  shorter  cut-off 
allows  of  greater  expansion  during  the  rolling  process.  The  crank  shaft  consists 
of  three  similar  pieces  coupled  by  flanges,  so  that  any  one  of  them  may  be  easily 
replaced  in  case  of  failure.  The  eccentric  shaft  is  carried  on  the  frame  and  is  driven 
from  the  crank  shaft  by  means  of  spur  gears.  This  allows  of  the  use  of  smaller 
eccentrics,  renders  the  valve  gear  more  accessible  on  top  of  the  engine  frame  and 
brings  the  center  lines  of  the  units  closer  together. 

The  piston  valves  are  arranged  on  top  of  the  cylinders  to  one  side  of  the  center 
lines  in  such  a  way  that  the  high  and  low  pressure  valves  of  one  tandem  unit  are 
in  line  and  are  both  driven  by  the  same  Stephenson  link  gear.  The  latter  has 

13* 


196 


crossed  eccentric  rods  as  is 
usual  in  rolling  mill  practice. 
A  stop  valve  is  provided  on 
each  of  the  six  cylinders, 
these  valves  being  operated 
by  an  auxiliary  power  cylinder 
which  at  the  same  time  con- 
trols the  position  of  the 
Stephenson  link  in  such  a 
manner  that  for  long  cut-offs 
the  steam  is  throttled,  while 
for  short  cut-offs  the  stop  val- 
ves are  fully  opened.  The  stop 
valves  of  the  low  pressure  cy- 
linders thus  allow  a  certain 
amount  of  steam  to  accumu- 
late in  the  receivers  when 
stopping  the  engine. 

The  problem  was  to  adapt 
this  very  satisfactory  design 
for -use  with  una-flow  cylinders 
while  retaining  as  far  as  pos- 
sible all  its  valuable  features. 
This  was  a  change  which  it 
would  pay  to  undertake,  since 
three  cylinders  with  their 
distance  pieces,  valve  gear 
and  accessories  could  be  dis- 
pensed with.  The  driving 
parts  of  course  must  be 
strengthened  to  take  the  higher 
pisto-n  loads  of  the  una-flow 
engine.  Such  a  design  by  Ehr- 
hardt  &  Sehmer  is  shown  in 
Figs.  6  and  7.  The  general 
structure,  i.  e.  frame,  valve 
gear,  and  the  use  of  piston 
valves  has  been  retained.  The 
triple  arrangement  of  this  en- 
gine, as  in  the  previous  case, 
will  also  give  the  advantage 
of  early  cut-offs  and  long  ex- 
pansions during  rolling.  To 
this  end  the  maximum  cut-off 
of  the  valve  gear  is  made 
somewhat  short,  but  in  order 


197 


198 


199 


200 


201 


202 


to  insure  plenty  of  power  for  starting,  the  piston  valve  bushing  contains  an 
auxiliary  port  which  gives  a  considerable  increase  of  cut-off  and  reduction  of  lap. 
This  arrangement  would,  however,  result  in  too  much  lead ;  and  in  order  to  prevent 
this  the  auxiliary  port  is  connected  to  the  cylinder  at  such  a  distance  from  the 

end  of  the  latter  that  at  the  time 
of  steam  admission  the  port  has 
been  overrun  by  the  piston  and  is 
straddled  by  a  pair  of  rings. 

If  for  instance  the  maximum 
cut-off  of  the  main  valve  is  35% 
and  the  cut-off  of  the  auxiliary 
port  70%,  then  at  slow  speeds 
shortly  after  starting,  or  when 
gripping  the  ingot,  the  cut-off  of 
70%  will  be  effective.  As  soon  as 
the  engine  comes  up  to  speed, 
however,  and  especially  when  it 
begins  to  race  after  the  passage  of 
the  ingot  through  the  rolls,  the 
auxiliary  port  cannot  supply  suffi- 
cient steam  to  make  itself  noticeable, 


bo 

£ 


and  the  cut-off  falls  to  practically  35%.  A  con- 
siderable saving  of  steam  and  increased  safety 
of  operation  are  the  result.  The  limitation  of  the 
maximum  cut-off  of  the  main  valve  has  the 
advantage  that  for  early  cut-offs  the  port  openings 
are  considerably  improved,  which  is  especially 
important  in  view  of  the  essentially  unfavorable 
valve  opening  consequent  on  the  use  of  crossed 
eccentric  rods. 

The  use  of  a  piston  valve  with  a  large  exhaust 
lap,  in  combination  with  the  link  valve  gear, 
permits  of  a  reduction  in  the  length  of  com- 
pression. The  same  purpose  is  served  by  an 
auxiliary  valve  mounted  in  the  main  piston 
valve,  by  means  of  which  the  compression  may 
be  reduced  or  almost  entirely  eliminated  by  an 
adjustment  made  from  the  operating  platform. 

Each  cylinder  is  fitted  with  a  separate  stop 
valve  for  throttling  or  cutting  off  the  steam 
supply  from  the  platform,  if  operating  conditions 
require  it. 

The  throttling  which  takes  place  when 
running  with  late  cut-offs  permits  of  gentle 
and  gradual  starting  of  the  engine.  The  steam 
must  in  fact  be  throttled  to  a  greater  extent  in 
a  una-flow  engine  owing  to  its  greater  starting 
torque  as  compared  with  that  of  a  tandem  com- 
pound. The  steam  consumption  during  starting 
will  therefore  be  correspondingly  less.  Even  with 
such  throttling,  the  early  maximum  cut-offs  used 
in  this  triple  engine  will  produce  appreciable 
expansion. 

If  the  ingot  should  stall  the  engine,  it  is  a 
simple  matter  to  exhaust  the  steam  within  the 
cylinders  by  reversing  the  gear;  and  on  again 
admitting  steam,  the  ingot  will  be  freed  from 
the  rolls.  The  above  described  valve  gear  will 
therefore  safeguard  the  operation  of  the  engine 
even  against  such  eventualities.  Separate  auxiliary 
exhaust  ports  are  not  necessary  since  the  inlet 
ports  are  used  for  this  purpose. 

The  piston  valves  are  preferably  designed 
with  inside  admission.  The  steam  space  between 
stop  valve  and  piston  valve  should  be  made  as 
small  as  possible.  A  comparison  between  this 
una-flow  rolling  mill  engine  and  a  tandem  com- 


203 


205 


fi 
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206 

pound  for  the  same  purpose  will  demonstrate  the  fact  that  considerable  simpli- 
fication, cheaper  construction  and  a  reduction  in  floor  space  are  attainable  with 
this  design.  Such  an  engine  is  also  essentially  more  powerful. 

The  best  engine  for  rolling  mill  service  is  the  one  which  consumes  the  least 
steam  during  the  comparatively  long  and  frequent  periods  of  idling.  The  claim 
made  by  Ehrhardt  &  Sehmer  that  the  una-flow  rolling  mill  engine  is  the  only  one 
which  satisfies  this  condition  is  justified,  since  it  alone  has  a  theoretically 
correct  no-load  diagram  and  the  least  no-load  steam  consumption  owing  to  the 
una-flow  exhaust,  to  the  long  compression  and  the  comparatively  large  inlet 
areas  in  consequence  of  the  rather  early  maximum  cut-off. 

In  compound  engines  the  steam  distribution  becomes  very  poor  when  using 
early  cut-offs  below  20%,  and  it  is  advisable  to  use  the  throttle  instead,  to  adjust 
the  output  to  the  load.  In  contrast  to  the  unfavorable  changes  of  exhaust  lead  and 
compression  in  the  compound  engine,  these  exhaust  phases  are  always  the  same  in 
the  una-flow  engine,  since  they  are  determined  by  the  exhaust  ports.  The  no-load 
diagram  must  accordingly  always  be  correct. 

In  Fig.  8  is  reproduced  a  series  of  continous  indicator  diagrams  taken  from 
a  flywheel  una-flow  rolling  mill  engine,  which  shows  clearly  the  no-load  cards  as 
well  as  the  rapid  succession  of  no-load  and  full  load  cut-offs,  thus  demonstrating 
the  excellent  governing  characteristics  of  the  engine.  This  of  course  applies  equally 
to  the  reversing  engine. 

A  very  interesting  design  of  a  flywheel  una-flow  rolling  mill  engine,  40  X  48". 
110  R.  p.  M.  max,  built  by  the  Mesta  Machine  Co.,  Pittsburgh  Pa.,  is  shown  in 
Fig.  9  and  10.  The  power  of  the  engine  is  transmitted  by  spur-gearing  on  the  roller 
shaft.  The  live  steam  enters  below  into  the  jackets  on  the  ends  of  the  cylinder 
barrel,  feeding  also  the  hollow  head  covers.  The  exhaust  belt  is  separated  by  two 
neutral  divisions  from  the  jacketed  ends  of  the  cylinder.  The  steam  enters  the 
cylinder  through  resilient  poppet  valves  placed  on  the  cylinder  barrel.  The  hollow 
piston  is  carried  by  a  heavy  hollow  piston  rod,  supported  by  the  crosshead  and  a 
slipper.  A  common  governor  controls  the  cut-off  by  shifting  a  small  crosshead  on 
the  operating  eccentric.  From  this  crosshead  the  motion  is  transfered  by  a  rod 
with  cam  and  roller  on  the  inlet  valves. 

The  whole  design  is  heavy,  strong,  reliable  and  especially  adapted  for  rolling 
mill  work,  and  all  precaution  is  taken  as  by  safety  valves,  draining  valves,  railing, 
stairs,  lagging,  enclosures  a.  s.  f.  for  securing  best  efficiency  and  maintenance  of 
the  engine. 

Attention  may  be  called  to  the  comparatively  flat  steam  consumption 
curve  (Fig.  9)  for  the  una-flow  engine,  which  not  only  shows  a  lower  steam 
consumption  than  the  compound  engine,  but  also  a  more  nearly  uniform  steam 
consumption  over  wide  ranges  of  load.  This  latter  feature  especially  recommends 
the  una-flow  engine  for  rolling  mill  service,  where  great  and  suddem  variations 
of  the  lead  are  the  rule. 


207 


Fig.  1. 


5.  The  Una-Flow  Hoisting  Engine. 

1.  For  Condensing  Service. 

The  una-flow  engine  is  suitable  for  use  as  a  hoisting  engine  if  means  are  pro- 
vided to  eliminate  the  compression  during  the  periods  of  starting  and  stopping, 
in  order  to  facilitate  the  exact  stoppage  of  the  cage.  An  auxiliary  valve  (see 
Fig.  1)  may  be  employed  for  this  purpose,  having  sufficient  lap  to  reduce  the  com- 
pression to  almost  nothing  when  starting 
or  when  the  cage  is  being  brought  into 
the  desired  position,  while  during  hoisting 
the  full  compression  of  about  90%  is 
effective.  During  the  hoisting  period  the 
auxiliary  valve  is  inoperative  as  regards 
the  compression  and  the  latter  is  deter- 
mined entirely  by  the  piston-controlled 
exhaust  ports  of  the  cylinder.  As  shown 
in  the  diagram  of  Fig.  2,  the  compression 
is  respectively  about  25,  65  and  90%  for 

cut-offs  of  80,  50  and  40%.  High  economy  in  the  utilization  of  steam  on  the  one 
hand,  and  excellent  maneuvering  abilities  on  the  other,  are  thus  combined  in  the 
best  possible  manner.  When  the  main  inlet  valves 
are  slide  or  piston  valves,  they  may  also  control 
the  auxiliary  exhaust,  and  the  exhaust  lap  should 
then  be  proportioned  from  the  above  standpoint. 

Figs.  3  and  4  show  a  design  recommended  for  a 
small  hoisting  engine.  Instead  of  the  auxiliary  piston 
valve  shown  in  Fig.  1,  two  poppet  valves  are  provided 
at  the  ends  of  the  cylinder,  which  come  into  ope- 
ration only  at  late  cut-offs.  This  arrangement  per- 
mits of  a  considerable  reduction  of  the  clearance 
volume  and  surfaces. 

The  valve  gear  is  of  the  Gooch  type  and  the 
motion  is  transmitted  to  both  the  inlet  and  auxiliary 
exhaust  valves  by  means  of  a  cam  mechanism.  The 
valve  diagrams  for  this  engine  are  similar  to  those 

shown  in  Fig.  2.  With  this  construction  it  becomes  possible  to  run  the  engine 
non-condensing  for  short  periods  if  the  exhaust  valves  and  exhaust  lap  are  pro- 
portioned accordingly.  For  this  temporary  condition  it  is  permissible  to  use 
rather  small  valves  with  correspondingly  high  steam  velocities. 

The  force  necessary  to  operate  this  gear  is  so  very  small  that  reversing  and 
notching  up  may  be  carried  out  by  hand,  thus  dispensing  with  a  power  maneuvering 


Fig.  2. 


208 

cylinder.  For  this  reason  it  is  advisable  to  use  the  Gooch  valve  gear,  in  which  only 
the  slide  block  and  valve  rod  have  to  be  lifted,  instead  of  the  link  and  eccentric  rods. 
The  four  valves  may  also  be  operated  by  means  of  the  ordinary  tapered  cam 
gear  (Figs.  5  and  6).  The  cams  are  preferably  arranged  so  as  to  give  only  a  small 
inlet  valve  lift  for  cut-offs  of  80  to  90%,  and  to  hold  open  the  comparatively 
small  auxiliary  valves  during  almost  the  whole  of  the  compression  stroke.  This 


Fig.  3. 


Fig.  4. 

enables  the  cage  to  be  stopped  with  accuracy  in  the  desired  position.  A  late  cut- 
off is  also  used  at  the  commencement  of  hoisting,  with  the  auxiliary  valves  like- 
wise in  action.  During  the  greater  part  of  the  hoisting  period  the  cut-off  is  early 
and  the  auxiliary  valves  are  out  of  action,  so  that  the  engine  operates  as  a  true 
una-flow  engine  under  the  most  favorable  conditions. 

The  tapered  cam  gear  may  also  be  operated  by  hand  in  most  cases  without 
the  use  of  an  auxiliary  power  cylinder.    Since  the  clearance  space  for  90%  com- 


209 

pression  may  be  properly  proportioned  to  give  sufficient  compression  for  any 
condenser  pressure,  the  maneuvering  of  the  engine  may  be  freely  and  easily  ac- 
complished without  depending  on  safety  devices  such  as  spring-loaded  valves  to 
prevent  excessive  compression. 

The  condenser  air  pump  should  be  preferably  independently  driven. 


Fig.  5. 


Fig.  6. 

2.  The  Una-flow  Hoisting  Engine,  Exhausting  to  Atmosphere  or  to  a 

Low  Pressure  Turbine. 

Fig.  7  shows  a  una-flow  hoisting  engine  for  non-condensing  service  built  by 
the  Gutehoffnungshiitte  Works  for  the  Vondern  Colliery,  shaft  No.  2.  The  engine 
is  designed  to  hoist  eight  tubs  of  500  kg  net  each,  from  a  depth  of  600  m,  with 


SLumpf,  The  una-flow  steam  engine. 


14 


210 

a  maximum  velocity  of  20  m/sec.  It  is  fitted  with  two  rope  drums  each  of  6400  mm 
diameter  and  1900  mm  width,  which  are  adjustably  coupled  for  hoisting  from 
different  levels.  The  engine  at  present  works  non-condensing,  with  steam  of  8  at. 
gauge  pressure.  Both  cylinders  have  a  bore  of  1100  mm  and  a  stroke  of  1600  mm. 
The  cylinders  are  unjacketed  plain  cylindrical  castings  supported  at  their  ends 
on  feet  resting  on  base  plates.  The  inlet  is  controlled  by  piston  valves  which  are 
arranged  to  give  a  supplementary  exhaust  while  the  cage  is  being  brought  to  rest 


Fig.  7. 

and  during  the  first'  few  strokes  after  starting.  The  compression  is  thus  almost 
entirely  relieved,  thereby  rendering  possible  an  exact  stoppage  of  the  cage  and 
easy  starting  of  the  engine. 

The  clearance  volume  is  10%  with  an  additional  clearance  pocket  of  8%. 
The  valves  are  indirectly  operated  by  tapered  cams  mounted  on  a  short  cross 
shaft  placed  at  right  angles  to  the  cylinder  at  its  middle  and  driven  from  the 
crank  shaft  by  means  of  bevel  gears  and  a  lay  shaft.  The  cams  act  on  pilot  valves 
which  control  auxiliary  pistons  coupled  to  the  main  piston  valves.  The  tapered 
cams  are  shifted  by  means  of  the  reversing  lever  without  the  use  of  a  power 
cylinder. 

This  hoisting  engine  is  equipped  with  a  combined  depth  gauge  and  safety 
device  of  the  Gutehoffnungshiitte  type  which  controls  the  hoisting  operation  from 


211 


14* 


212 

beginning  to  end.  It  prevents  overspeeding  and  slows  the  engine  down  automatically 
and  in  a  predetermined  manner  when  the  cage  approaches  its  stopping  place. 
The  safety  device  not  only  controls  the  cut-off  as  the  engine  gets  under  way,  but 
also  causes  the  steam-operated  brake  to  come  into  action  gradually,  according  to 
the  amount  of  overspeed,  and  to  release  the  brake  when  the  speed  again  falls. 
If  the  cage  passes  its  stopping  point,  the  brake  is  applied  with  full  power.  In- 
creased safety  is  thus  imparted  to  both  hoisting  and  stopping. 

The  engine  has  given  complete  satisfaction  as  regards  service,  but  not  in 
respect  to  steam  consumption,  and  this  is  fully  accounted  for  by  the  low  initial 
pressure,  the  large  clearance  volume  and  the  use  of  piston  valves. 

The  una-flow  engine  is  particularly  well  suited  for  the  work  of  a  hoisting 
engine  on  account  of  the  direct  action  of  the  steam.  The  greater  the  number  of 
expansion  stages,  the  more  sluggish  the  engine  will  be;  and  conversely,  the  smaller 
the  number  of  stages  the  more  lively  it  will  be  in  its  action.  For  this  reason,  and 
also  on  account  of  the  higher  diagram  factor,  the  stroke  volume  of  the  una-flow 
cylinder  can  be  made  considerably  smaller  than  that  of  the  low  pressure  cylinder 
of  a  compound  hoisting  engine,  both  in  respect  to  stopping  the  cage  and  for  the 
accelerating  period.  In  the  latter  connection  it  may  be  mentioned  that  the  una- 
flow  engine  will  run  at  higher  mean  effective  pressures  with  the  same  steam  con- 
sumption as  the  compound  engine. 

With  the  una-flow  engine  there  is  no  need  to  worry  over  the  maintenance  of 
the  compound  effect  for  the  various  changes  of  load,  or  when  the  engine  is  tem- 
porarily at  rest.  The  radiation  losses  will  also  be  considerably  smaller  as  com- 
pared with  those  of  the  four  cylinders  and  receivers  of  a  twin  tandem  compound. 
These  radiation  losses  are  especially  large  in  the  latter  since  the  receiver  pressure 
must  be  maintained  while  the  engine  is  temporarily  at  rest.  On  the  other  hand, 
the  una-flow  hoisting  engine  may  use  more  steam  for  maneuvering. 

Apart  from  its  thermal  superiority,  the  una-flow  hoisting  engine  has  obvious 
constructional  advantages.  In  order  to  make  these  clear,  a  comparison  has  been 
made  in  Fig.  8  between  a  twin  tandem  compound  hoisting  engine  of  usual  design 
having  cylinders  of  900  and  1400  mm  bore  by  1800  mm  stroke,  and  a  una-flow 
hoisting  engine  of  the  same  power,  the  cylinders  of  which  would  have  a  diameter 
of  about  1250  mm.  The  length  of  the  una-flow  cylinder  casting  would  be  3000  mm 
as  against  2900  of  the  low  pressure  cylinder.  The  overall  length  of  the  una-flow 
engine  would  be  6  m  less  than  the  length  of  the  tandem  compound  engine.  The 
engine  house  and  the  foundation  would  be  shorter  by  the  same  amount.  Two 
complete  cylinders  with  valve  gear,  two  distance  pieces  and  two  receivers  are  dis- 
pensed with.  The  oil  consumption  will  be  correspondingly  smaller  and  the  whole 
engine  will  be  cheaper,  simpler  and  more  reliable,  notwithstanding  its  heavier 
driving  parts. 


213 


6.  Una-Flow  Engines  for  driving  Air  Compressors, 

Pumps,  etc. 

The  constructional  simplification  of  the  una-flow  engine  is  of  particular 
advantage  in  connection  with  air  compressors,  pumps  or  blowing  tubs.  Two- 
stage  engines  are  usually  built  as  cross-compounds,  which  works  out  satisfactorily 
in  many  cases.  The  una-flow  engine,  however,  permits  of  a  straight  line  con- 
struction; and  although  this  can  also  be  employed  in  two-stage  engines  in  a  tan- 
dem arrangement,  it  is  somewhat  inconvenient  and  has,  therefore,  not  found 
much  favor. 

The  una-flow  engine  is  of  course  also  suitable  for  a  twin'  arrangement  as 
shown  in  Fig.  1.  This  illustration  represents  a  una-flow  pumping  engine  built 
by  the  firm  of  Gustav  List  in  Moscow  for  a  municipality  in  Central  Russia.  The 
point  of  importance  in  this  case  was  reserve  power,  with  first  cost  as  a  somewhat 
secondary  consideration.  This  requirement  is  satisfactorily  met  in  this  engine, 
since  each  side  is  a  complete  unit  and  can  be  operated  independently,  although 
with  a  different  flywheel  effect.  A  surface  condenser  is  arranged  crosswise  under- 
neath the  cylinders  and  either  of  the  latter  may  be  blanked  off  from  it  by  means 
of  a  slip  flange,  inserted  between  the  connecting  flanges.  Each  cylinder  is  also 
provided  with  an  independent  change-over  valve  and  exhaust  pipe  for  non-con- 
densing operation.  The  pumps  are  placed  in  a  well  and  driven  from  the  engine 
tail  rod  by  means  of  a  bell-crank  which  also  drives  the  condenser  air  pumps.  The 
surface  condenser  is  cooled  by  the  water  delivered  by  the  main  pumps,  so  that 
no  circulating  pumps  are  necessary. 

In  Fig.  2  is  shown  an  air  compressor  in  which  the  air  and  steam  cylinders 
are  combined.  The  steam  cylinder  is  single-acting  and  the  inlet  valve  is  of  the 
single-beat  type  forged  in  one  piece  with  the  stem.  The  valve  is  actuated  by  a 
rolling  lever  mechanism  enclosed  in  a  housing  and  running  in  oil.  This  mechanism 
comprises  a  rocking  spindle  operated  by  the  eccentric,  which  spindle  carries  a  curved 
lever  acting  upon  one  arm  of  a  rolling  lever,  which  in  turn  opens  the  valve. 

By  keeping  the  whole  mechanism  running  in  oil,  the  wear  of  the  moving  parts 
is  almost  entirely  eliminated.  The  clearance  volume  of  the  steam  end  is  ex- 
tremely small,  since  the  valve  stem  may  be  brought  close  to  the  cylinder  barrel.  An 
auxiliary  exhaust  valve  is  provided  to  reduce  the  compression.  In  order  to  shorten 
the  cylinder  as  much  as  possible,  an  exhaust  lead  of  20°/0  is  used  and  the  port 
area  is  correspondingly  reduced. 

This  larg ;  exhaust  lead  permits  of  the  use  of  a  partial  exhaust  belt  and  allows 
the  cylinder  and  piston  to  be  shortened.  The  air  end  is  of  standard  design.  The 
suction  valves  consist  of  a  split  spring  steel  band  or  ring  covering  the  suction 
ports  which  are  drilled  in  the  cylinder  flange,  and  the  cylinder  head  forms  the 
valve  guard.  The  whole  cylinder  head  surface  is  available  for  the  arrangement 
of  the  discharge  valves,  which  is  of  particular  advantage  in  small  compressors. 


Fig.  1. 


216 


Fig.  3. 


Fig.  5. 


217 


so 

£ 


218 

The  discharge  valve  is  also  in  the  form  of  a  split  spring  steel  band  mounted  in  the 
cylinder  head  so  as  to  cover  the  delivery  holes  communicating  with  the  cylinder. 
The  clearance  volume  of  the  air  end  also  is  very  small.  Very  favorable  steam 
consumption  results  may  be  expected  with  this  construction,  in  view  of  the  ex- 
cellent test  results  obtained  with  a  similar  cylinder  design  given  in  the  chapter 
on  locomobiles. 

The  cut-off  may  be  changed  by  buckling  the  eccentric  rod  by  means  of  a 
handwheel  and  link. 

The  central  exhaust  ports  half  way  between  the  steam  and  air  ends,  with 
the  outlet  at  the  lowest  point  of  the  cylinder,  will  make  it  impossible  for  water 
to  get  into  the  compressor  side.  Simplicity,  low  first  cost  and  accessibility  of  all 
important  parts  are  advantages  of  this  design. 

One  such  engine  as  shown  in  Fig.  3  has  been  built  by  the  firm  of  A.  L.  G. 
Dehne,  of  Halle  a.  d.  S.  For  larger  units  a  two-crank  arrangement  would  offer 
certain  advantages,  one  side  comprising  a  standard  horizontal  or  vertical  una- 
flow  engine  and  the  other  a  standard  horizontal  or  vertical  blowing  tub,  air  com- 
pressor, etc.  It  would  be  desirable  in  this  case  to  arrange  the  cranks  in  such  ,a 
way  as  to  reduce  the  flywheel  weight  to  a  minimum. 

Fig.  4  shows  a  una-flow  driven  straight  line,  two-stage  air  compressor  built 
by  the  Linke- Hoffmann  Works,  of  Breslau.  The  cylinders -are  arranged  in  tandem, 
the  air  cylinder  being  next  to  the  frame,  with  a  distance  piece  between  the  air 
and  steam  cylinders.  The  differential  air  piston  at  the  same  time  performs  the 
function  of  a  crosshead.  The  high  pressure  stage  is  at  the  crank  end,  and  the  low 
pressure  stage  at  the  head  end  of  the  air  piston.  The  intercooler  is  placed  in  the 
foundation  below  the  air  cylinder.  The  steam  valves  of  the  una-flow  cylinder 
are  operated  by  a  Stumpf  gear  and  a  governor  on  the  crank  shaft.  Auxiliary 
exhaust  valves  are  also  provided,  likewise  actuated  by  a  cam  mechanism  driven 
from  the  crankshaft,  by  means  of  which  the  length  of  compression  may  be  reduced 
to  about  one-half  of  that  given  by  the  central  exhaust  ports.  The  engine  operates 
condensing  without  the  use  of  the  auxiliary  exhaust  valves,  and  their  valve  gear 
should  therefore  be  arranged  so  that  it  may  be  disconnected. 

A  high  speed  pumping  engine  built  by  the  Worthington  Pump  &  Machinery 
Corporation,  of  New  York  City,  is  shown  in  Figs.  5,  6  and  7.  In  this  case  the  una- 
flow  cylinder  is  placed  next  to  the  frame;  and  the  double-acting  pump,  arranged 
in  tandem  with  it,  is  connected  to  the  steam  end  by  tie  rods  running  from  the 
pump  body  to  the  crank  end  cylinder  head.  The  steam  cylinder  bore  is  13%", 
pump  plunger  diameter  11",  stroke  21",  speed  210  r.  p.  m.,  steam  pressure 
220  Ibs/sqin.  gauge,  steam  temperature  562.4°  F.  and  total  head  including  suction 
lift  153  ft.  The  steam  valves  are  horizontal  and  are  actuated  by  tapered  cams 
on  the  lay-shaft,  which  runs  on  ball  bearings  and  is  driven  by  spiral  gears  from 
the  crank  shaft.  The  section  of  the  lay-shaft  carrying  the  cams  is  shifted  endwise 
by  a  speed  governor  so  as  to  control  the  maximum  speed.  A  pressure  regulator 
acting  in  a  similar  manner  is  mounted  at  the  end  of  the  lay-shaft  and  governs  the 
engine  for  constant  water  pressure.  The  pump  is  provided  with  nests  of  auto- 
matic valves  of  special  design,  which  are  accessible  from  b'oth  sides  by  means 
of  hinged  manhole  covers.  An  interesting  feature  is  the  complete  enclosure  of 


219 


220 


t>0 

£ 


221 


Fig.  8. 


222 


Fig  9. 


223 

all  working  parts,  which  even  extends  to  the  lay-shaft,  governor,  piston  rod,  crank 
shaft  and  flywheel,  so  that  no  moving  part  is  visible.  The  automatic  lubricating 
system  includes  all  bearings,  pins  and  glands.  The  engine  has  proved  very  satis- 
factory in  service. 

In  Fig.  8  and  9  is  shown  the  application  of  a  una-flow  cylinder  to  a  recipro- 
cating tube  pump,  an  indicator  card  of  which  is  reproduced  in  Fig.  10.  This  type 
of  pump  is  being  developed  by  the  Humphrey  Gas  Pump  Company,  of  Syracuse, 
N.  Y.,  and  consists  of  a  tube  provided  with  a  foot  valve,  which  is  reciprocated  in 
a  well  or  casing  by  direct  attachment  to  the  piston  of  a  una-flow  cylinder  mounted 
at  the  well  head.  The  weight  and  energy  of  the  tube  and  its  contents  are  utilized 
in  connection  with  the  functions  of  the  power  medium,  which  in  this  machine  is 
steam,  by  permitting  the  latter  to  be  used  expansively.  This  is  in  marked  con- 
trast to  the  usual  type  of  direct- 
acting  well  pump,  where  late 
cut-offs  are  necessary,  thus  re- 
sulting in  excessive  steam  con- 
sumption. 

The  cylinder  has  a  bore  of 
12",  with  a  stroke  of  approxi- 
mately 10",  the  speed  being 
from  150  to  200  cycles  per  mi- 
nute. The  cylinder  is  of  course  Fig.  10. 
single-acting  and  its  upper  end 

is  closed  to  form  a  cushion  chamber  which  serves  to  retard  the  reciprocating  parts 
and  tube  towards  the  end  of  the  upstroke  when  the  exhaust  ports  are  uncovered, 
while  the  water  continues  to  move  through  the  tube.  Hence  practically  the  whole 
of  the  kinetic  energy  of  the  moving  parts  is  stored,  and  thus  becomes  available  for 
accelerating  them  on  the  downstroke.  This  energy,  as  well  as  that  due  to  the  fall 
of  the  tube,  is  then  used  in  the  compression  of  the  residual  steam  in  accordance  with 
the  una-flow  cycle.  While  this  is  taking  place  the  water  is  still  in  motion  relatively 
to  the  tube,  and  this  continues  until  admission  begins  at  the  commencement  of 
the  upstroke,  when  the  tube  is  again  accelerated  and  the  foot  valve  closes. 

The  valve  gear  consists  of  a  swinging  link  and  lever  operated  from  the  cross- 
head  to  which  the  rods  carrying  the  tube  are  attached.  Since  the  length  of  the 
stroke  is  not  positively  determined  by  mechanical  means,  a  delayed  action  of  the 
inlet  valve  is  provided  for  by  causing  the  valve  gear  lever  to  operate  a  pilot  valve, 
which  in  turn  controls  the  admission  and  release  of  steam  behind  a  piston  forming 
part  of  the  double-beat  inlet  valve.  The  latter  is  of  cast  iron  and  works  on  a  resi- 
lient lower  steel  seat.  The  valve  is  cushioned  both  on  opening  and  closing  by  a 
double-acting  adjustable  oil  dash  pot,  which  is  insulated  as  far  as  possible  from 
the  valve  chest  cover.  This  valve  gear  has  proved  satisfactory  and  quiet  in  action, 
but  contains  a  somewhat  large  number  of  parts.  A  direct  acting  piston  valve 
mechanism  has  also  been  built.  The  volumetric  efficiency  is  considerably  over  100%. 

A  great  una-flow  blowing  engine,  built  by  the  Mesta  Machine  Co.,  Pittsburgh 
Pa,  is  shown  in  Fig.  11. 


224 


Fig.  11.     Una-Flow  Blowing  Engine  built  by  the  Mesta  Machine  Company  Pittsburgh. 


225 


III.  The  Una-Flow  Locomotive. 

Fig.  1  shows  a  superheater  freight  engine  which  was  built  for  the  Moscow- 
Kasan  Railway  by  the  Kolomna  Engine  Works,  of  Kolomna,  near  Moscow. 
This  was  the  first  locomotive  built  upon  recommendation  of  Mr.  Noltein,  the 
manager  of  the  railway,  and  was  fitted  with  una-flow  cylinders  designed  by  the 
author.  They  were  originally  fitted  with  small  auxiliary  exhaust  valves  which 
were  removed  later,  so  that  the  engine  now  operates  as  a  true  una-flow. 


Fig.  1. 


Fig.  2. 

• 

Fig.  2  shows  one  of  a  pair  of  superheater  freight  locomotives  built  for  the 
Prussian  State  Railways  by  the  Stettiner  Maschinenbau  -  Aktiengesellschaft 
,,Vulkan".  In  Fig.  3  is  given  a  longitudinal  and  cross  section  of  the  locomotive, 
as  well  as  a  longitudinal  section  of  the  cylinder.  Figs.  4  and  5  show  the  simpli- 
city of  the  cylinder  and  valve  gear  parts.  All  the  experience  obtained  with  the 
above-mentioned  Russian  locomotive  was  utilized  in  this  design.  These  engines 
were  put  in  heavy  continuous  day  and  night  service  and  proved  so  successful  that 
a  number  of  engines  of  the  same  design  were  ordered. 

Slumpf,  The  una-flow  steam  ensine.  15 


226 

The  valves  and  live  steam  spaces  are  arranged  in  the  heads  while  the  exhaust 
ports  and  exhaust  chamber  or  belt  are  at  the  middle  of  the  cylinder.  This 
strict  separation  of  hot  live  steam  from  the  cold  exhaust  steam  is  not  only 
advisable  for  thermal  reasons  but  also  from  an  operating  standpoint,  since  the 
parts  exposed  to  high  temperatures  cannot  affect  the  operating  conditions  of 


-136O 


Fig.  3. 


the  piston  or  cause  warping  of  the  cylinder,  and  the  central  exhaust  belt 
effectively  cools  the  middle  part  of  the  latter  where  the  piston  attains  its 
highest  velocity. 

For  the  non-condensing  service  of  locomotives  it  was  necessary  to  provide 
a  large  clearance  volume,  in  this  case  of  17%%,  for  superheated  steam  of  12  at. 
This  large  clearance  is  the  weak  point  of  the  design,  and  in  later  engines 


227 


it  was  materially  reduced.  The  greater  part  of  the  clearance  space  is  disposed 
in  the  concave  ends  of  the  piston  (Fig.  6).  The  piston  heads  thus  take  the  shape 
of  spherical  caps,  great  strength  and  stiffness  being  thereby  obtained.  They  are 
made  of  cast  steel  and  are  fitted  with  two  piston  rings  each.  A  distance  piece  or 
drum,  made  of  hard  forged  steel  7  mm  in  thickness,  is  placed  between  the  heads 
and  the  whole  is  clamped  together  by  means  of  the  piston  rod  and  nut,  for  which 
purpose  a  certain  amount  of  clearance  is  left  between  the  hubs  of  the  piston  heads. 
The  supporting  drum  has  a  clearance  of  three  thousandths  of  the  diameter,  so 
that  for  a  cylinder  bore  of  1000  mm  its  diameter  would  be  997  mm.  This  would 
bring  the  piston  center  line  1  %  mm  below  the  cylinder  center.  The  allowance 
takes  care  of  the  expansion  and  distortion  of  the  cylinder  and  drum.  Special  con- 
sideration must  be  paid  to  the  expansion  of  the  heads,  since  they  are  exposed 
to  the  full  live  steam  tem- 
perature during  admission. 
They  therefore  expand  more 
than  the  supporting  drum  and 
thus  distend  the  ends  of  the 
latter,  whereby  scoring  of  the 
cylinder  may  be  caused.  This 
was  a  source  of  difficulty  in 
some  of  the  earlier  pistons, 
but  was  overcome  by  using 
a  sphere  of  smaller  radius  for 
the  piston  heads  and  provi- 
ding a  larger  allowance  for  the 
ends  of  the  supporting  drum. 

Lubrication  is  effected, 
not  by  introducing  oil  into 
the  steam  chest,  but  by 
feeding  it  to  a  number  of 
points  in  the  cylinder  wall, 
and  thus  bringing  it  directly 
to  the  working  surfaces.  Each 
end  of  the  cylinder  has  three  Fig.  4. 

oil  feeds,  one  on  top  and  one 

on  each  side  on  the  horizontal  center  line,  each  feed  being  supplied  by  an 
independent  plunger.  Even  with  this  arrangement  the  oil  may  still  carbonize 
and  it  is  therefore  better  to  place  the  feeds  closer  to  the  middle  of  the 
cylinder. 

A  tail  rod  with  its  attendant  stuffing  box  was  not  used  on  these  locomotives. 
A  gain  in  weight  thus  results,  but  the  long  piston,  the  length  of  which  is 
given  by  the  stroke  less  the  exhaust  lead,  generally  works  out  heavier  than 
the  standard  piston  with  tail  rod. 

The  necessary  clearance  volume  depends  upon  the  pressure  and  temperature 
of  the  live  steam.  Assuming  the  back  pressure  to  be  1,1  at.  abs.  (0,1  at.  being  added 
for  the  resistance  in  the  blast  pipe),  90%  length  of  compression  (adiabatic),  quality 

15* 


228 

of  steam  =  1,  and  terminal  compression  pressure  =  live  steam  pressure,  then  for 
different  steam  pressures  the  following  clearances  are  necessary: 

Steam  pressure,  at.  gage     11       12       13       14       15       16       17       18       19 
Clearance  volume    .    .  %  16.9    15.8    14.6    13.9    13.1    12.6    12.1    11.6    11.1 

In  using  this  table  it  must  be  borne  in  mind  that  a  pressure  loss  of  about 
1  at.  occurs  in  the  superheater,  and  that  the  clearance  should  be  increased  by 
such  an  amount  that  for  normal  cut-off  the  difference  between  the  terminal  corn- 


Fig.  5. 

pression  and  live  steam  pressure  is  equal  to  the  difference  between  the  terminal 
expansion  and  back  pressure,  according  to  the  rules  given  in  the  chapter  on  vo- 
lume loss. 

The  una-flow  action  is  not  limited  to  the  steam,  but  applies  also  to  any 
foreign  matter  contained  in  it,  such  as  scale,  mud,  cinders  or  soot.  The  latter 
are  swept  out  by  the  exhaust  steam  through  the  central  ports  and  escape  through 
an  orifice  or  drain  at  the  lowest  point.  The  una-flow  action  therefore  permanently 
maintains  the  interior  of  the  cylinder  in  a  clean  state. 

The  drain  just  mentioned  also  insures  the  removal  of  water,  and  thus  elimi- 
nates a  difficulty  occurring  in  all  ordinary  locomotives.  The  cylinder  in  the  latter 
forms  the  lowest  point  of  the  system,  the  live  steam  enters  from  the  top,  and  the 
exhaust  steam  leaves  the  cylinder  also  at  the  top.  Damage  due  to  water,  such  as 
fractured  cylinders  and  covers,  as  well  as  breakage  of  driving  parts,  are  possible 
with  this  arrangement.  Nothing  of  this  kind  can  happen  with  a  una-flow  locomo- 
tive, since  the  water  is  effectively  cleared  from  the  cylinder  by  the  exhaust  steam 


229 

and  passes  away  through  the  drain.  It  is  surprising  to  see  how  much  water  is  ejected 
through  this  drain  when  starting  with  a  cold  cylinder.  A  different  kind  of  water 
hammer  may  be  caused  by  the  kinetic  energy  of  water,  apart  from  its  being  trapped 
in  the  cylinder,  and  this  may  of  course  happen  in  a  una-flow  engine. 

Fig.  3  also  shows  a  mechanism  by  means  of  which  both  valves  may  be 
lifted  off  their  seats  from  the  cab,  thus  putting  the  two  sides  of  the  piston 
in  communication  with  each  other  through  the  inlet  pipe,  and  relieving  the 
compression  when  coasting.  It  is  advisable  to  provide  considerable  lift  for  the 


Fig.  6. 


valves,  as  otherwise  when  running  down  a  long  grade  the  temperature  of  the 
steam  pulsating  to  and  fro  between  the  cylinder  ends  will  become  so  high  in 
consequence  of  friction  and  wiredrawing  that  the  oil  may  carbonize  and  cause 
piston  troubles.  The  admission  of  cool  air  to  the  cylinder  through  a  valve  opened 
simultaneously  with  the  lifting  of  the  steam  valves  is  also  recommended.  When 
the  steam  valves  are  lifted  together  with  their  cams,  the  rollers  should  run  clear 
of  the  latter  so  as  to  prevent  unnecessary  wear.  In  contrast  to  the  special  by-pass 
arrangements  used  in  ordinary  piston  valve  cylinders,  the  valve  gear  of  the  una- 
flow  locomotive,  with  only  minor  additions,  provides  a  by-pass  for  coasting 
without  increasing  the  clearance  volume  and  harmful  surfaces. 

The  exhaust  lead  is  usually  taken  at  10%,  thus  fixing  the  length  of  com- 
pression at  90%.    A  rapid  uncovering  of  too  large  a  port  area  may  produce  an 


230 


abrupt  exhaust  and  a  pulsating  draft  in  the  fire  box,  whereas  a  uniform  draft 
is  desirable  for  good  combustion.  This  may  be  realized  by  using  a  large  exhaust 
lead  in  combination  with  a  correspondingly  small  area  of  ports  and  of  the  blast 
pipe,  as  well  as  by  interposing  some  form  of  receiver  or  exhaust  chamber,  thus 
allowing  of  better  expansion  of  the  exhaust  and  serving  to  muffle  the  noise. 

In  ordinary  locomotives  the  steam  distribution  for  very  early  cut-offs  is  un- 
satisfactory, and  for  this  reason  throttling  of  the  steam  is  usually  employed  instead 
of  making  the  cut-off* early er.  In  contrast  to  this,  the  una-flow  locomotive  can 
be  run  entirely  without  throttling,  and  allows  the  power  requirements  to  be  regu- 
lated entirely  by  means  of  the  valve  gear.  The  constant  compression  on  the  other 
hand  is  a  disadvantage,  particularly  in  regard  to  the  large  clearance  volume,  and 

makes  itself  especially  felt 
when  running  with  large 
cut-offs. 

Fig.  7  shows  a  resilient 
valve  made  of  forged  steel, 
for  a  passenger  locomotive. 
The  valves  are  operated 
by  a  cam  and  roller  mecha- 
nism similar  to  that  used 
with  the  Stumpf  valve 
gear  for  stationary  engines 
(Fig.  8).  The  cam  rollers 
are  carried  in  milled  grooves 
in  the  reciprocating  slides, 
the  grooves  at  the  same 
time  serving  as  oil  retainers. 
The  guides  are  long  enough 
to  prevent  the  grooves  from 
overrunning  them,  thereby 
Fig.  7.  preventing  loss  of  oil  and 

the  entrance  of  dust.     The 

oil  lubricating  the  cam  crosshead  or  guide  collects  mostly  in  these  grooves,  and 
is  transferred  by  the  rollers  to  the  cams  so  that  proper  lubrication  of  these  impor- 
tant parts  is  insured.  The  guides  themselves  are  provided  with  wick  oilers.  The 
roller  slides  are  operated  by  a  standard  Walschaert  gear  without  any  changes 
from  that  used  on  existing  counterflow  locomotives.  Both  roller  slides  have  screw 
adjustments. 

The  valve  spring  is  arranged  to  be  adjustable  and  must  be  powerful  enough 
to  allow  of  running  with  late  cut-offs  at  high  speeds. 

Fig.  9  shows  a  double-seated  automatic  compression  release  valve  which  is 
in  communication  with  both  ends  of  the  cylinder  through  two  pipe  connections. 
The  entrance  to  each  pipe  is  controlled  by  a  valve  which  may  be  operated  from 
the  cab.  When  these  valves  are  opened  the  live  steam  from  one  cylinder  end  closes 
the  corresponding  side  of  the  compression  release  valve  and  opens  the  other,  so 
that  the  compression  steam  of  the  opposite  cylinder  end  may  escape  through  the 


231 

passage  between  the  two  seats  of  the  auxiliary  valve.  When  admission  occurs 
at  the  opposite  end  of  the  cylinder,  the  auxiliary  valve  changes  its  position,  so 
that  compression  is  now  relieved  on  the  other  side.  The  passage  between  the 


seats  of  the  valve  may  be  connected  with  the  exhaust  belt.  This  device  therefore 
allows  compression  to  be  entirely  eliminated,  so  that  a  great  reserve  of  power 
becomes  available  for  the  difficult  starting  period.  The  shut-off  valves  at  the  cylin- 


232 

ders  are  closed  as  soon  as  the  train  is  in  motion,  and  the  release  valve  thus  be- 
comes inoperative.  Experience  has  shown,  however,  that  the  device  is  not  abso- 
lutely essential. 

Fig.  10  illustrates  a  una-flow  freight  locomotive  exhibited  at  the  Brussels 
International  Exposition  of  1910,  cross-sections  of  which  are  shown  in  Fig.  3. 
In  order  to  compensate  for  the  increased  weight  of  the  una-flow  cylinders,  the 
boiler  has  been  moved  back  on  the  frame  by  a  small  amount  for  better  distri- 
bution of  the  axle  loads. 


Cylinder  bore 600  mm 

Stroke 660    „ 

Driving  wheel  dia.      .    .  1350    ,, 
Steam  pressure,  gauge   .        12  at. 

Boiler  heating  surface    .  140.42  sqm 


Superheater  surface       .    .        38.97  sqm 

Grate  area      2.35     ,, 

Weight,  empty 52125kg 

57750   , 


Service  weight 


Fig.  9. 


Fig.  11    shows   a   una-flow   freight   locomotive   built   by   the  Schweizerische 
Lokomotivfabrik,  of  Winterthur,  for  the  Swiss  Federal  Railways. 


Cylinder  bore 570  mm 

Stroke 640    „ 

Driving  wheel  dia.      .    .  1330    „ 

Leading  truck  wheel  dia.  850    „ 

Steam  pressure,   gauge  .  12  at. 


Boiler  heating  surface    .    143.7  sqm 
Superheater  surface    .    .     37.6     „ 

Grate  area 2.44 

Weight,  empty    .    .    .    .60875    kg 
Service  weight     ....  67  700 


The  general  design  of  the  cylinders  is  similar  to  those  previously  described, 
with  the  exception  that  a  muffler  has  been  incorporated  in  the  saddle  supporting 
the  smoke  box,  between  the  exhaust  belt  and  the  blast  pipe  (Fig.  12). 

Fig.  14  shows  a  una-flow  passenger  locomotive  built  by  the  Maschinenbau- 
anstalt  Breslau  for  the  German  State  Railways.  An  engine  of  this  type  was 
exhibited  at  the  Turin  Exposition. 


233 


234 


Cylinder  bore   .    .    .    .    .        550  mm 

Stroke    . 630    „ 

Steam  pressure,  gauge  .          12  at. 

Boiler  heating  surface     .  136.98  sqm 

Superheater  surface    .    .  40.32     „ 

Total  heating  surface     .  177.30     „ 


Driving  wheel  dia. 
Truck  wheel  dia. 
Grate  area  .  .  . 
Weight,  empty  . 
Service  weight  . 
Adhesive  weight 


2100  mm 
1000    „ 
2.31  sqm 
56900  kg 
62500     „ 
35000 


The  Northern  Railway  of  France  had  a  superheater  freight  locomotive 
fitted  with  una-flow  cylinders  which  were  designed  by  the  author  in  a  similar 
way  to  those  previously  described,  the  existing  Stephenson  link  valve  motion 
being  retained. 

In  Fig.  14  is  shown  a  una-flow  passenger  locomotive  for  saturated  steam,  two 
of  which  were  built  by  the  Kolomna  Engine  Works  for  the  Russian  State  Rail- 
ways. Two  other  engines  of  the  same  design  were  fitted  with  superheaters.  The 
former  run  with  14,  and  the  latter  with  12  at.  gage  pressure.  The  extra  pressure 


Fig.  12. 

carried  by  the  engines  using  saturated  steam  was  rendered  possible  for  the  same 
axle  loads  by  putting  the  weight  of  the  superheater  into  the  heavier  boiler  plates. 
Since  there  is  a  loss  of  about  1  at.  in  the  superheater,  there  is  a  gain  of  3  at.  in 
favor  of  the  locomotive  using  saturated  steam.  All  previous  experiences  and  test 
results  were  utilized  in  the  design  of  the  cylinders.  Attention  is  directed  especially 
to  the  careful  jacketing  of  the  heads  and  ends  of  the  cylinders  for  saturated  steam. 
The  condensate  from  the  jackets  is  returned  to  the  boiler  by  a  small  pump  with 
suction  ports,  operated  by  a  cam  on  the  roller  rod. 


Main  dimensions  of  the  engines  \ising  saturated  steam: 


Cylinder  bore 500  mm 

Driving  wheel  dia 1700     „ 

Steam  pressure,  gauge    .    .        14  at. 


Stroke  .... 
Heating  .surface 
Grate  area 


650  mm 
166.57  sqm 
2.45 


A  series  of  comparative  tests  were  made  by  Prof.  Lomonosoff  on  the 
Russian  State  Railways  on  one  each  of  the  saturated  and  superheater  una- 
flow  locomotives,  in  competition  with  two  ordinary  saturated  steam  com- 


2—6—0  Una-Flow  Locomoti 


f  the  German  State  Railways. 


235 


236 


pounds   and  one  superheater  compound  locomotive, 
are  as  follows: 


The  results  of  these  tests 


2  —  6  —  0  two  cylinder  loco- 

Saturated  Steam 

Superheated  Steam 

1700mm  Three-axle  tender 

Compound 

Una-flow 

Una-flow 

Compound 

Train  weight  350  tons  I 
Tver  to  Moscow       j 

V 
G 

51.5 
35.6 

50.3 
37.7 

50.5 
35.7 

54.3 
30.8 

49.6 
29.2 

Z 

42.8 

36.9 

41.5 

45.8 

37.4 

Train  weight  270  tons  I 
Moscow  to  Tver       | 

V 
G 

67.3 
43.9 

71.0 
38.8 

73.3 
32.7 

76.3 
29.4 

69.0 
32.0 

Z 

41.1 

42.1 

41.3 

42.5 

38.1 

Train  weight  424  tons  I 
Moscow  to  Tula      1 

V 
G 

49.4 
28.3 

51.6 
31.8 

43.3 
35.3 

47.9 
25.0 

44.6 
23.7 

Z 

39.8 

42.9 

49.5 

42.5 

34.7 

Train  weight  270  tons  J 
Tula  to  Moscow       j 

V 
G 

56.8 
40.1 

56.8 
38.5 

55.4 
36.4 

59.9 
28.0 

57.9 
30.1 

Z 

40.5 

36.0 

34.5 

37.5 

33.7      » 

V  =  Mean  speed  in  km/hour. 

G  ==  Naphta  consumption  per  1  ton-km. 

Z  =  Evaporation  per  1  sqm  of  boiler  heating  surface. 

These  figures  permit  of  a  fair  comparison  of  the  economy  of  the  locomotives; 
since  for  one  run  the  total  number  of  ton-km  is  the  same  for  all  locomotives,  their 

output  must  also  be  practi- 
caly  equal,  except  for  changes 
in  the  train  resistance  caused 
by  wind  and  other  weather 
conditions.  The  evaporation 
Z  per  1  sqm,  however,  is 
different  for  every  locomotive 
and  no  conclusion  can  there- 
fore be  drawn  from  the  same. 
The  line  from  Moscow 
to  Tver  is  almost  level,  but 
that  from  Moscow  to  Tula 
has  long  and  heavy  grades 
in  both  directions. 

The  conclusions  which  can  be  drawn  from  these  tests  are  as  follows: 
The  una-flow  locomotive  shows  better  economy  than  the  compound  for  small 
loads,  while  at  higher  loads  its  fuel  consumption  is  higher  than  that  of  the  latter. 
This  can  be  easily  explained  by  the  effect  of  the  long  constant  compression  and 
the  large  clearance  volume  (see  chapter  on  volume  loss). 

The  una-flow  locomotive  working  with  saturated  steam  shows  in  general  a 
higher  economy  than  the  compound  except  for  long  cut-offs.  Larger  cylinders 
would  be  of  advantage  in  this  case. 


Fig.  15. 


0—8—0  Una-Flow  Locomoti 


:  the  Moscow  Narrow  Gage  Railway. 


16. 


237 


The  superheater  una-flow  locomotive  is  at  least  on  a  par  with  the  super- 
heater compound,  although  even  here  the  former  has  a  slightly  higher  fuel  con- 
sumption for  heavy  loads.  Larger  cylinders  are  of  course  more  feasible  with  the 
una-flow  system  than  with  the  compound  engine. 


Fig.  15  shows  a  una-flow  cylinder  with 
horizontal  valves  for  a  small  locomotive 
built  by  the  Kolomna  Engine  Works.  The 
mechanism  operating  these  valves  is  of  in- 
terest; it  consists  of  a  double  armed  lever, 
the  lower  end  of  which  works  against  the 
valve  stems  while  its  upper  end  receives 
its  motion  from  a  rocking  lever  fitted  with 
two  cam  profiles  which  are  alternately  in 
rolling  contact  with  it.  The  rocking  lever 
is  driven  by  a  Marshall  gear.  Attention  is 
drawn  to  the  accessibility  of  the  valves 
which  are  removable  after  taking  off  the 
valve  chest  covers,  without  disturbing  any 
other  part  of  the  gear. 

Fig.  17.  It  will  be  noted  that  the  Kolomna  En- 

gine Works  consistently  adhere  in  all  their 

designs  to  the  true  una-flow  arrangement.    Much  value  is  placed  by  these  builders 
on  the  series  arrangement  of  live  steam  space,  inlet  valve,  piston  and  exhaust. 


238 


In  Fig.  17  are  given  the  main  sections  of  a  una-flow  cylinder  for  a  narrow 
gage  locomotive  (Fig.  16)  using  saturated  steam,  built  by  the  Kolomna  Engine 
Works,  which  was  shown  at  the  Turin  Exposition.  In  view  of  the  small  size  of 
the  cylinders,  the  latter  are  fitted  with  slide  valves,  which  are  separate  for  each 
cylinder  end.  Since  these  valves  are  of  small  area  compared  with  that  of  the 
ordinary  D- valve,  the  load  upon  them,  as  well  as  the  resistance  which  the  valve 
gear  has  to  overcome,  is  considerably  diminished.  Reliable  operation  is  insured 
by  feeding  oil  under  pressure  to  their  working  faces. 


Cylinder  bore 355  mm 

Stroke 350    „ 

Driving  wheel  dia.      .    .  750    ,, 

Steam  pressure    ....  12  at.  gage 


Boiler  heating  surface 

Grate  area 

Weight  empty      .    .    . 
Service  weight     .    .    . 


54.26  sqm 
0.93     „ 
19.2  tons 
21.2 


Fig.  18. 


Fig.  19. 


Fig.  20. 


It  is  the  policy  of  the  Kolomna  Engine  Works,  in  cases  where  superheating 
is  not  acceptable,  to  obtain  improved  economy  by  applying  the  una-flow  system. 
In  this  way  they  rebuilt  with  una-flow  cylinders  eleven  saturated  steam  loco- 
motives of  the  tank  type  of  the  Warsaw  narrow  gage  railway  during  the  years 
1913 — 1914.  These  una-flow  cylinders  are  fitted  with  piston  valves  and  have  the 
cylinder  ends  jacketed,  in  addition  to  the  heads. 


239 

The  use  of  two  rows  of  central  exhaust  ports  with  a  simultaneous  increase  of 
the  exhaust  lead  from  10  to  30%  considerably  shortens  the  cylinder  and  piston, 
as  shown  in  Fig.  18.  The  second  set  of  exhaust  ports  provides  a  very  effective 
increase  in  port  area  shortly  before  dead  center.  If  this  advantage  is  not  con- 
sidered essential,  then  the  use  of  a  single  row  of  ports  with  the  same  exhaust  lead 
of  30%  will  still  further  reduce  the  length  of  the  cylinder  and  piston  (Fig.  19). 

It  is  important  to  reduce  the  port  area  as  the  exhaust  lead  is  lengthened  in 
order  to  keep  the  loss  of  diagram  area  at  the  end  of  the  stroke  within  reasonable 
limits.  (See  diagram  Fig.  20.)  The  long  exhaust  lead  at  the  same  time  shortens 


Fig.  21. 

the  compression  from  90%  to  70%  and  reduces  the  clearance  volume  from  16,2% 
to  13,6%.  This  reduction  of  the  length  of  the  cylinder  and  the  saving  in  weight 
of  both  cylinder  and  piston  should  prove  of  value  particularly  for  locomotives. 
The  design  of  the  una-flow  locomotive  cylinder  shown  in  Fig.  21  differs  from 
those  previously  described  in  that  the  piston  valve  gives-a  supplementary  exhaust 
for  long  cut-offs  and  the  engine  operates  on  the  true  una-flow  cycle  only  for  early 
cut-offs.  As  is  shown  by  the  indicator  diagrams,  the  true  una-flow  cycle  is  main- 
tained for  cut-offs  up  to  30%.  For  longer  cut-offs  an  auxiliary  (counter- flow) 
exhaust  comes  into  effect,  which  shortens  the  compression  with  increasing  cut-off. 
In  this  way  starting  is  facilitated  for  long  cut-offs,  and  the  advantage  of  the  full 
una-flow  cycle  for  steady  running  is  yet  retained.  The  compression  release  device 
shown  in  Fig.  9  is  thus  dispensed  with  in  this  case.  The  varying  length  of  the 
compression  considerably  reduces  the  volume  loss  at  late  cut-offs,  but  this  is  ob- 
tained at  the  expense  of  the  series  arrangement  of  live  steam,  inlet  valve,  piston  and 


240 

exhaust.  The  piston  valve  has  inside  steam  admission,  and  the  outside  lap  con- 
trols the  supplementary  exhaust.  The  inlet  and  exhaust  at  each  end  of  the  piston 
valve  are  in  parallel  with  the  piston,  and  steam  leaking  by  the  valve  will  pass 
directly  into  the  exhaust.  The  supplementary  exhaust  is  conducted  from  the 
ends  of  the  piston  valve  housing  to  the  exhaust  belt  through  separate  pipe  connec- 
tions. The  piston  is  built  up  of  two  cast  steel  pieces,  each  of  which  is  fitted  with 
a  bronze  shoe.  Fig.  22  shows  a  similar  piston  valve  design  which  was  used  for 
two  passenger  locomotives  of  the  German  State  Railways.  A  corresponding 
design  was  employed  for  the  cylinders  of  a  una-flow  locomotive  of  the  North  Eastern 
Railway  of  England  (Fig.  23),  and  for  those  of  the  locomotives  of  the  Neuruppin- 
Kremmen-Wittstock  Railway  (Fig.  24). 

This  same  cylinder  and  piston  valve  design  was  also  used  for  the  three  cylinder 
passenger  superheater  locomotives  (Fig.  25)  built  for  the  German  State  Rail- 
ways by  the  Vulkan  Engine  Works  of  Stettin.  The  three  cranks  are  set  at  120°. 


Fig.  22. 

The  two  outside  cylinders  drive  the  second  coupled  axle,  and  the  inside  cylinder, 
which  is  placed  slightly  ahead  of  the  others,  operates  the  first  driving  axle.  By 
distributing  the  piston  loads  upon  two  axles  a  longer  life  of  the  center  crank  axle 
is  expected,  especially  since  the  forging  can  be  made  with  easy  curves,  without 
disturbing  the  natural  fibers  of  the  material.  The  motion  of  the  valve  of  the  inside 
cylinder  is  derived  from  the  two  outside  Walschaert  gears,  the  movement  of  which 
is  combined  according  to  the  parallelogram  of  velocities.  The  head  ends  as  well 
as  the  crank  ends  of  the  three  cylinders  may  be  connected  by  means  of  special 
valves  in  order  to  relieve  compression  when  coasting. 

An  examination  of  the  exhaust  timing  of  a  three  cylinder  locomotive  shows 
that  the  exhaust  periods  of  the  individual  cylinders  overlap;  a  very  uniform  draft 
in  the  fire  box  is  thus  obtained,  particularly  if  an  auxiliary  exhaust  is  provided 
for  long  cut-offs.  The  three  cylinders  constitute  an  excellent  reinforcement  of  the 
locomotive  frame.  The  clearance  volume  of  each  cylinder  is  11%,  the  bore  500  mm 
and  the  stroke  630  mm. 


241 


Fig.  23. 


Fig.  24. 


Stumpf,  The  una-flow  steam  engine. 


Fig.  25. 


16 


242 


fcC 

£ 


243 

The  piston  valve  bushings  are  designed  in  such  a  way  as  to  prevent  catching 
of  the  rings  when  removing  or  reassembling  the  valves.  The  pistons  are  fitted 
with  bronze  shoes  in  the  middle  of  their  length  so  as  to  avoid  contact  with  the 
hot  part  of  the  cylinder. 

Noteworthy  is  the  small  clearance  volume  of  only  11%,  as  well  as  the  manner 
of  arranging  the  clearance  spaces  so  as  to  reduce  the  harmful  surfaces  to  a  mini- 


Fig.  27. 

mum.  For  this  reason  the  piston  valve  is  made  single-ported  instead  of  the  more 
common  double-ported  construction,  and  this  is  compensated  for  by  the  long 
travel  of  190  mm. 

The  locomotives  are  equipped  with  superheaters  of  the  Schmidt  type  and 
exhaust  steam  feed  water  heaters. 

In  the  first  locomotives  described  in  this  chapter,  the  true  una-flow  principle 
was  adhered  to  by  the  author  in  order  to  obtain  a  minimum  surface  loss  in  addition 
to  the  other  advantages  mentioned.  This  minimum  surface  loss,  however,  is 
accompanied  by  a  large  volume  loss.  When  working  with  superheated  steam  a  small 

16* 


244 


J610 


Fig.  28. 


245 

surface  loss  can  also  be  realized  in  the  counterflow  construction  if  the  whole  cycle 
takes  place  in  the  superheated  region,  as  is  nearly  always  the  case  in  modern 
superheater  locomotives.  A  simultaneous  reduction  of  the  surface  and  volume 
losses  to  a  minimum  is  possible  with  una-flow  locomotives  in  the  following  manner. 
A  reduction  of  the  volume  loss  while  retaining  the  full  una-flow  cycle  is  pos- 
sible by  raising  the  initial,  or  lowering  the  back  pressure.  The  latter  way  avoids 
the  difficulties  arising  from  a  considerable  increase  in  the  boiler  pressure  and  is 
based  on  the  utilization  of  the  large  amount  of  energy  still  contained  in  the  toe 
of  the  diagram,  for  the  purpose  of  reducing  the  back  pressure.  This  principle  was 
described  in  chapter  I,  7.  Its  first  application  is  found  on  a  superheater 
freight  locomotive  of  the  German  State  Railways,  built  in  1920  by  A.  Borsig 
of  Berlin,  according  to  designs  furnished  by  the  author.  The  main  dimensions 
of  the  locomotive,  which  is  illustrated  in  Figs.  26  to  28,  are  as  follows: 


Cylinder  bore 630  mm 

Stroke 660   „ 

Driving  wheel  dia.      .    .  1400    ,, 

Maximum  speed  ....  60  km/hour 

Steam  pressure    ....  12  at.  gage 

Grate  area 2.62  sqm 


Boiler  heating  surface    .    .  149.65  sqm 

Superheater  heating  surface  53.00    ,, 

Total  heating  surface    .    .  202.65    „ 

Feed  water  heater  surface  15.0    ,, 

Weight  empty    .    .    .    .    ?  65.5  tons 

Service  weight 72.0   „ 


This  brings  the  una-flow  locomotive  into  a  new  phase  of  development,  since 
the  lower  back  pressure  reduces  the  compression  pressures  and  permits  of  the 
use  of  smaller  clearance  volumes.  The  exhaust  ejector  action  also  produces  an 
approximately  correct  variation  of  the  compression  with  the  cut-off,  since  the 
energy  available  in  the  large  toe  of  late  cut-off  cards  produces  a  strong  ejector 
effect,  with  a  correspondingly  low  back  pressure  and  a  low  compression  pressure; 
while  at  early  cut-offs  the  ejector  action  is  less  pronounced  and  the  back  pressure 
and  terminal  compression  pressure  are  higher.  In  order  to  obtain  the  exhaust 
ejector  effect  a  large  exhaust  lead  is  essential,  and  the  latter  at  the  same  time 
shortens  piston  and  cylinder.  With  this  long  duration  of  the  exhaust  only  a  small 
port  area  is  required,  with  the  result  that  the  exhaust  belt  can  be  dispensed  with 
and  a  considerable  reduction  in  the  weight  of  the  cylinder  and  piston  thus  results. 
A  comparison  of  Figs.  27  and  28  with  the  previous  designs  indicates  how  much 
more  compact  this  construction  has  become.  This  is  in  part  due  to  the  use  of 
horizontal  single-beat  poppet  valves  which  were  employed  for  the  first  time  on 
this  locomotive.  This  type  of  valve,  although  simple  and  perfectly  steam  tight, 
has  so  far  not  been  favorably  received  because  it  requires  a  high  lift  and  a  large 
force  to  raise  it.  With  the  high  compression  of  the  una-flow  engine,  however,  the 
pressure  on  the  valve  is  balanced  to  a  large  extent  and  the  high  lift  is  obtained 
by  arranging  the  cam  roller  between  the  valve  stem  and  the  fulcrum  of  the  valve 
lever.  The  lift  of  the  cam,  which  is  14  mm  radially,  is  thus  increased  to  24  mm 
at  the  valve.  For  cut-offs  up  to  50%  the  effective  inlet  areas  of  the  single-beat 
valve  are  equivalent  to  the  areas  of  a  standard  piston  valve  of  220  mm  diameter. 
The  fact  that  beyond  this  cut-off  the  valve  area  remains  constant  must  be  consi- 
dered a  further  advantage.  The  small  cam  lift  permits  of  a  cam  profile  of  very 
gentle  curvature,  thus  insuring  smooth  lifting  and  seating  of  the  valve.  The  whole 


247 


cam  mechanism  is  very  substantially  constructed  and  swinging  levers  were  used 
instead  of  sliding  parts  wherever  possible.  It  should  therefore  stand  up  well  in 
service. 

The  single  beat  valve  is  made  of  chrome-nickel  steel  and  works  on  a  remo- 
vable steel  seat  expanded  into  the  cylinder  casting  (Fig.  30).  If  this  seat  should 
become  damaged  by  scale  or  other  foreign  matter  it  can  be  easily  resurfaced  or 
renewed.  The  valve  stem  has  a  diameter  of  25  mm  and  is  supplied  with  oil  under 
pressure.  The  common  center  of  gravity  of  the  valve  head  and  spring  retainer 
is  located  at  about  the  center  of  the  guide  so  that  good  working  conditions  are 
assured.  The  valve  stem  furthermore  is  not  exposed  to  the  live  steam  but  to  the 
varying  pressure  and  temperature  of  the  cylinder  steam.  It  is  entirely  independent 
of  the  cam  mechanism  except  for  the  tappet  contact,  and  is  free  to  follow  any  slight 
distortion  of  the  cylinder  casting.  Considering  the  success  of  the  horizontal  valves 
in  the  Lanz  locomobile,  which  are 
double-beat  in  addition,  the  con- 
clusion is  justified  that  this  is  a 
very  reliable  construction. 

When  coasting,  the  valves  may 
be  lifted  off  their  seats  by  com- 
pressed air  admitted  between  small 
pistons  formed  on  the  valve  tappets, 
so  that  the  rollers  clear  the  cam. 
Special  means  fcr  connecting  the 
cylinder  ends  are  therefore  not  re- 
quired, and  the  usual  relief  valves 
may  be  omitted,  since"  the  inlet 
valves  act  as  such.  They  also  relieve 
the  high  compression  which  may  occur  when  the  throttle  is  nearly  closed.  The 
automatic  compression  release  device  also  may  become  superfluous  since  the 
late  cut-offs  at  starting  produce  a  strong  exhaust  ejector  effect  and  the  com- 
pression is  therefore  considerably  shortened. 

Attention  may  be  drawn  to  the  accessibility  of  the  valves;  for  their  renewal 
it  is  only  necessary  to  take  off  the  valve  chest  cover  and  disconnect  the  valve 
spring,  the  spring  cap  lock  being  a  split  spherical  washer.  Comparing  this  with 
the  procedure  of  taking  out  an  ordinary  piston  valve,  which  requires  frequent 
removal  of  carbonized  oil,  the  great  simplification  due  to  the  single  beat  valve 
will  be  appreciated. 

The  driving  parts  and  the  Walschaert  gear  are  the  same  as  those  used  on 
counterflow  locomotives.  On  account  of  its  greater  length  the  cylinder  was  moved 
forward  by  180  mm.  The  una-flow  cylinder  is  not  heavier  than  the  corresponding 
counterflow  cylinder,  since  the  piston  valve  chest  with  its  large  exhaust  chamber, 
as  well  as  the  tail  rod  and  its  guide  are  omitted.  This  allows  the  piston  rod  of 
95  mm  diameter  to  be  bored  out  to  a  diameter  of  60  mm,  thus  also  saving  weight. 
The  forged  steel  piston  heads,  which  are  only  slightly  dished,  hold  between  them 
a  cast  iron  supporting  drum  cast  from  a  special  soft  mixture,  while  the  cylinder 
is  made  of  a  hard  quality  of  cast  iron.  The  supporting  drum  is  turned  smaller 


Fig.  30. 


248 


than  the  cylinder  bore  by  2,2  mm  on  a  length  of  140  mm  at  its  middle,  which 
allowance  increases  to  5  mm  towards  the  ends.  Each  piston  head  carries  three  rings. 
The  greater  part  of  the  total  clearance  volume  of  12%  is  taken  up  by  a  linear 
clearance  of  40  mm  between  piston  and  cylinder  head,  and  this  also  results  in  very 
small  harmful  surfaces.  The  pressure  oil  feeds  are  arranged  at  the  middle  of  the 
cylinder,  where  the  temperature  is  lowest  and  little  possibility  of  carbonizing  exists. 
One  feed  is  placed  on  top  and  one  on  each  side  at  45°  below  the  horizontal  center 
line. 

It  may  be  a  matter  of  surprise  that  hardly  anything  is  ever  heard  of  attempts 
to  utilize  the  energy  of  the  exhaust  steam  of  one  cylinder  to  produce  a  vacuum 

in  another.  Such  experiments  have  been 
made,  for  instance  in  connection  with 
Kordina's  blast  pipe,  but  they  were 
bound  to  fail  in  counterflow  locomotives. 
As  was  shown  in  chapter  I,  7,  in  dealing 
with  the  exhaust  ejector  effect,  a  large 
part  of  the  available  energy  is  require'd 
for  producing  the  draft  in  Ihe  fire  box, 
and  if  the  energy  is  used  in  a  wasteful 
fashion  there  will  be  nothing  left  for  the 
reduction  of  back  pressure.  The  ordi- 
nary counterflow  cylinder  has  exhaust 
passages  of  such  an  uneven  character 
that  the  steam  continuously  suffers 
changes  of  direction  and  velocity  which 
naturally  dissipate  part  of  its  energy. 
This  is.  well  shown  by  a  comparison  of 
Fig.  31  with  Fig.  28.  The  una-flow  cy- 
linder is  designed  on  the  principle  of 
conserving  the  exhaust  energy,  while  in 
the  counterflow  design  it  might  be 
thought  that  a  dissipation  of  energy 
was  aimed  at.  However,  no  blame  for 
this  can  be  laid  on'  the  designer,  since  tortuous  and  uneven  exhaust  passages 
are  inseparable  from  the  use  of  piston  valves. 

In  consequence  of  these  conditions,  a  certain  pressure  difference  at  the  blast 
nozzle  becomes  necessary,  since  velocity  must  again  be  generated.  It  may  be  con- 
sidered an  excellent  result  if  a  pressure  difference  of  0,1  at.  gage  at  the  blast  nozzle 
is  found  sufficient.  In  many  cases,  however,  the  kinetic  energy  is  dissipated  to 
such  an  extent  that  there  is  not  enough  left  to  cover  the  blast  pipe  loss.  The  latter 
can  only  be  diminished  by  a  reduction  of  the  velocity.  Since  the  blast  pipe  area 
is  fixed,  this  reduction  of  velocity  can  only  be  accomplished  by  a  diminution  of  the 
volume,  which  thus  leads  to  an  increase  of  pressure.  In  this  way  the  back  pres- 
sure in  piston  valve  cylinders  occasionally  rises  to  0,5  at.  In  contrast  to  this  con- 
dition, a  back  pressure  of  0,5  at.  below  atmosphere  is  aimed  at  with  the  ejector 
effect  of  the  una-flow  exhaust  for  late  cut-offs. 


Fig.  31. 


249 

Although  the  possibility  of  the  utilization  of  the  exhaust  energy  for  the  pur- 
pase  of  reducing  the  back  pressure  has  no  inherent  connection  with  the  una-flow 
principle,  it  is  limited  to  the  latter  by  virtue  of  the  conditions  of  exhaust  whereby 
the  energy  is  conserved,  and  it  must  therefore  be  considered  an  integral  part  of 
the  same. 

The  exhaust  opening  in  the  cylinder  wall  is  in  the  shape  of  a  nozzle  which 
is  dimensioned  in  such  a  way  that  in  combination  with  the  rounded  edge  of  the 
piston  an  approximately  correct  nozzle  area  for  any  particular  pressure  drop  is 
obtained,  when  assuming  average  pressure  and  speed  conditions.  The  remaining 
pressure  energy  is  therefore  transformed  into  kinetic  energy  and  thus  reaches  the 
blast  nozzle  with  the  least  possible  friction  losses.  Even  for  full  opening  there 
is  a  certain  amount  of  divergence  in  the  exhaust  nozzle.  At  the  junction  of  the 


I 

I 

E 

]3mi 

6.8* 

i,« 

2.  »* 

Fig.  32. 

two  exhaust  pipes  the  section  suddenly  doubles,  and  from  this  point  on  the  pipe 
diverges  to  the  blast  nozzle  and  thus  acts  as  a  kind  of  diffusor.  The  steam  for 
the  feed  water  heater  is  withdrawn  from  the  cylinders  by  separate  nozzles,  and 
the  pipes  from  them  lead  to  a  common  ejector  nozzle  similarly  to  the  main  ex- 
haust pipes. 

It  was  of  course  important  to  make  the  blast  pipe  section  as  large  as  pos- 
sible in  order  to  reduce  the  blast  pipe  loss.  The  stack  was  therefore  designed  with 
the  most  favorable  dimensions,  its  diameters  being  460  and  610  mm  as  compared 
with  410  and  460  mm  of  that  of  the  standard  engine.  The  possibility  therefore 
arose  that  the  jet  leaving  the  blast  nozzle,  which  was  dimensioned  to  obtain  a 
diffusor  action,  would  not  spread  out  sufficiently  to  fill  out  the  stack  section. 
This  danger  is  especially  great  in  this  case,  since  the  jet  has  very  little  internal 
pressure  and  spreads  out  to  only  a  comparatively  small  extent.  If  the  stack  is 
not  completely  filled  with  steam,  air  enters  the  smoke  box  from  above  and  thus 


250 


partially  destroys  the  vacuum.  In  order  to  prevent  this,  the  jet  may  be  divided 
by  ribs  in  the  nozzle  into  smaller  diverging  jets,  which  spread  out  and  again 
join  in  the  stack  (Fig.  28).  In  actual  operation  however  the  ribs  proved  to 
be  superfluous. 

The  effect  of  the  ejector  action  is  shown  in  the  diagrams  of  Fig.  32,  which 
were  drawn  for  an  evaporation  of  7000  kg/hour  and  for  speeds  of  20,  40  and 
60  km/hour,  under  very  conservative  assumptions.  The  clearance  volume  was 
assumed  to  be  12%,  so  that  the  compression  would  not  exceed  the  initial  pressure 
even  without  the  exhaust  ejector  action.  It  will  be  seen  from  the  diagram  that 
the  ejector  effect  only  begins  after  a  piston  travel  of  6,7%,  since  in  a  two  cylinder 
locomotive  the  cylinders  are  in  connection  only  during  a  short  part  of  the  stroke, 
and  therefore  only  part  of  the  exhaust  energy  can  be  utilized  for  producing  the 
ejector  effect.  It  was  shown  in  chapter  I,  7,  how  much  greater  the  gain  would 
be  with  a  three  cylinder  locomotive,  especially  when  working  with  high  mean 
effective  pressures,  since  in  this  case  with  proper  length  of  exhaust  lead  the  whole 
of  the  energy  of  the  diagram  toe  can  be  utilized  for  the  exhaust  ejector  action. 

In  connection  with  this,  it  may  be  mentioned  that  the  trend  of  modern  loco- 
motive design  is  toward  an  increasing  adoption  of  the  three  cylinder  locomotive. 

The  exhaust  ejector  principle,  however,  is  only  one  of  the  means  for  reducing 
the  volume  loss,  its  effect  being  the  lowering  of  the  back  pressure.  The  next  step 
will  consist  in  raising  the  upper  limit  of  the  steam  pressure.  With  the  customary 
design  of  fire  box,  steam  pressures  up  to  16  at.  gage  are  possible,  although  the 
number  and  size  of  the  stays  becomes  excessive.  For  still  higher  pressures  a  diffe- 
rent design  of  fire-box  would  be  necessary,  such  as  for  instance  a  box  of  the  Brotan 
type,  which  permits  of  steam  pressures  of  20  at.  gauge.  The  following  table  shows 
that  by  raising  the  steam  pressure  from  12  to  20  at.  gage,  the  amount  of  heat 
which  can  be  converted  into  work  increases  from  116  to  146  cal.  per  1  kg  steam. 
This  represents  a  gain  of  about  26%. 


Boiler 
pressure 

at  gauge 

Live  steam 

Exhaust  steam 

Heat 
con- 
verted 
into 
work 

Cal. 

Satur- 
ation 
point 

at.  abs. 

Clear- 
ance 
volume 

Cylinder 
pressure 

at  gauee 

Temp, 
at 
Cylinder 
°C 

Total 
heat 

Cal. 

Back 
pressure 

at.  abs. 

Total 
heat 

Cal. 

Counter-flow 

12 

11 

320 

740 

1,15 

624 

116 

2,2 

10 

Una-flow 

20 

19 

320 

734 

0.85 

588 

146 

4.75 

7 

This  remarkable  result  can  only  be  attained  by  employing  the  una-flow  prin- 
ciple, since  the  pressure  at  which  the  steam  becomes  saturated  increases  from 
2.2  to  4.75  at.  abs.  and  during  a  considerable  part  of  the  expansion  moisture  is 
therefore  formed  in  the  cylinder.  This  does  not  have  much  effect  upon  the  eco- 
nomy of  the  una-flow  engine  but  is  very  detrimental  to  that  of  the  counterflow 
cylinder  where  the  presence  of  moisture  causes  large  surface  losses. 

The  practice  with  counterflow  locomotives  is  therefore  to  use  higher  superheat 
with  higher  initial  pressure.  The  increase  in  temperature,  however,  is  the  cause 


251 

of  many  difficulties  with  piston  valves  and  piston  rod  packings.  Furthermore, 
the  superheater  elements  must  be  shortened  so  that  the  flue  gases  do  not  exert 
a  cooling  effect  upon  the  superheated  steam.  This  in  turn  leads  to  a  great  number 
of  superheater  elements  and  inefficient  utilization  of  space.  The  una-flow  engine 
can  of  course  be  adapted  to  meet  this  condition  in  a  better  manner,  since  its  design 
makes  it  more  suitable  for  high  temperatures  than  the  customary  counterflow 
cylinder  with  piston  valves.  On  the  other  hand  there  is  no  necessity  for  using 
these  high  temperatures  in  the  high  pressure  una-flow  locomotive,  since  the  una- 
flow  action  corrects  the  bad  influence  of  moisture  in  the  steam. 

The  calculated  gain  of  26%  of  the  high  pressure  una-flow  locomotive  with 
exhaust  ejector  action  will  probably  be  exceeded  in  practice,  since  it  does  not 
include  the  benefit  due  to  the  single  beat  poppet  valves,  the  reduction  of  the 
clearance  volume  to  7%,  nor  even  the  gain  due  to  the  una-flow  principle  itself. 

The  future  line  of  progress  of  the  locomotive  is  therefore  clear.  It  leads  natu- 
rally from  the  two-cylinder  to  the  three-cylinder  engine  with  una-flow  cylinders 
having  small  clearance  volumes,  to  the  use  of  single-beat  poppet  valves  and  the 
utilization  of  the  ejector  action  of  the  exhaust,  in  combination  with  high  pressures 
and  high  superheat. 


252 


IV.  The  Una-Flow  Locomobile  and  Portable  Engine. 

The  una-flow  engine  is  .especially  suitable  for  locomobiles  and  portable  engines. 
Simplicity,  lightness,  cheapness  and  economical  use  of  steam  are  demanded  of  this 
type  of  engine,  light  weight  being  particularly  required  for  portable  and  self  pro- 
pelled machines.  All  these  requirements  are  satisfied  by  the  una-flow  engine. 

Figs.  1  to  3  illustrate  the  construction  employed  by  the  Maschinenfabrik  Ba- 
denia  vorm.  Wm.  Platz  Sohne,  A.-G.,  Weinheim  (Baden). 

In  comparison  with  the  usual  design  of  locomobile  with  a  tandem  engine, 
there  is  a  saving  due  to  the  omission  of  one  cylinder  with  all  its  accessories.  Com- 
pared with  the  usual  type  of  compound  poppet  valve  locomobile,  one  complete 
driving  unit  and  two  valves  on  the  remaining  cylinder  are  dispensed  with.  While 
the  una-flow  locomobile  engine  has  only  two  valves,  the  compound  locomobile 
requires  eight.  The  omission  of  exhaust  valves  is  of  particular  advantage.  The 
cylinder  may  therefore  be  mounted  directly  on  the  boiler,  and  the  two  inlet  valves 
may  be  arranged  vertically  in  the  heads  (Fig.  4).  These  valves  are  operated  by 
an  eccentric  mounted  on  the  crank  shaft  next  to  the  flywheel,  controlled  by  the 
flywheel  governor.  The  motion  is  transmitted  to  the  valves  in  the  same  manner 
as  on  stationary  una-flow  engines.  Bolted  to  the  cylinder  is  a  frame  of  the  forked 
type  which  carries  the  center  crank  shaft.  On  the  free  ends  of  the  shaft  are  mounted 
two  overhung  flywheels,  one  of  which  carries  the  governor  and  the  other  is  pro- 
vided with  teeth  for  the  barring  device  (Fig.  5).  Provision  is  made  in  the  design 
for  interchanging  the  two  flywheels  so  that  the  condenser  may  be  placed  either 
to  the  right  or  the  left  of  the  engine.  All  the  parts  concerned  are  constructed  in 
such  a  way  as  to  make  them  suitable  for  either  method  of  setting  up,  and  a  good 
basis  for  large  scale  production  is  therefore  obtained. 

The  air  pump  is  placed  vertically  and  is  driven  from  a  crank  pin  on  the 
flywheel. 

Una-flow  locomobiles  are  built  for  condensing  service  as  well  as  atmospheric 
exhaust.  In  the  former  case,  non-condensing  operation  is  provided  for  by  the 
employment  of  auxiliary  clearance  pockets. 

The  reciprocating  masses  are  partially  balanced  by  counterweights  on  the 
crank  cheeks  and  in  the  flywheels. 

The  governor  shown  in  Figs.  6  and  7  has  two  weights  pivoted  on  two  pins 
fixed  on  the  flywheel.  These  weights  are  in  the  form  of  bell  crank  levers,  the 
short  arms  of  which  carry  rollers  bearing  on  the  thrust  washer  of  a  central  spring 
common  to  both.  The  long  arm  of  each  weight  is  provided  with  a  pin  at  its  extreme 
end,  but  only  one  of  them  is  used  for  the  connection  to  the  shifting  eccentric. 
The  free  weight,  however,  also  takes  part  in  moving  the  eccentric  to  the  extent 
that  it  bears  more  strongly  on  the  spring  plate  and  thus  relieves  the  second  weight, 


253 


Fig.  l. 


Fig.  2. 


254 


Fig.  4. 


255 

which  therefore  has  a  larger  force  available  for  shifting  the  eccentric.  The  primary 
eccentric  is  provided  with  tapped  holes  so  that  it  may  be  turned  through  an  angle 
of  180°  —  2  6  (6  being  the  angle  of  advance),  and  bolted  in  this  position.  If 
the  shifting  eccentric  is  at  the  same  time  swung  around  and  its  connecting  link 
attached  to  the  other  weight,  the  governor  is  then  changed  over  for  a  left  hand 
engine.  This  change  can  therefore  be  made  without  any  alteration  in  the  parts 
concerned.  This  governor  is  obviously  only  suited  for  high  speeds,  since  the 
natural  oscillations  due  to  the  weights  themselves  would  be  detrimental  to  regu- 
lation at  low  speeds. 

This  type  of  governor,  although  now  no  longer  used,  is  mentioned  here  in 
order  to  demonstrate  the  conditions  which  must  be  satisfied  by  a  locomobile  governor 
from  the  point  of  view  of  its  suitability  for  a  variety  of  uses. 

The  movement  of  the  shifting  eccentric  is  transmitted  to  the  valves  by  a 
rocking  lever  and  cam  and  roller  motion  in  the  manner  previously  described.  It 
is  advisable  to  make  the  rocking  lever  in  one  piece  of  cast  steel. 

In  Fig.  8  is  shown  the  assembly  of  a  100  HP  locomobile,  also  built  by  the 
Badenia  Company.  The  cylinder  is  bolted  to  the  boiler  and  is  in  rigid  connection 
with  the  main  bearings  through  the  forked  frame.  The  bearing  housings  rest  on 
half-round  pieces  of  steel,  so  that  free  expansion  of  the  boiler  and  correct  align- 
ment of  the  shaft  are  assured.  This  feature  has  been  patented  by  the  builders. 
The  bearings  are  of  double  construction,  so  that  the  inner  halves  on  the  one 
hand  closely  support  the  crank  .arms  and  take  the  steam  load,  while  the  outer 
halves  carry  the  flywheel  load.  Chain  lubrication  is  provided  for  in  the  center 
of  each  double  bearing.  One  of  the  flywheels  is  cast  with  teeth  for  barring  pur- 
poses and  the  other  carries  the  shaft  governor  described  above.  The  additional 
clearance  pockets  which  are  provided  to  allow  the  engine  to  be  run  non-condensing 
are  arranged  in  the  cylinder  heads.  The  exhaust  belt  of  the  cylinder  has  an  ex- 
tension forming  a  base  which  is  bolted  to  the  boiler.  The  exhaust  belt  has  an 
opening  on  either  side  for  the  connection  to  the  condenser.  The  flywheels,  frame, 
bearings,  cylinders  and  heads,  condenser,  all  the  valve  gear  parts,  governor,  barring 
device,  feed  water  heater,  change-over  valves,  and  all  accessories  are  constructed 
in  such  a  manner  as  to  allow  of  either  a  right  or  left  hand  arrangement,  to  provide 
for  condensing  or  non-condensing  service,  and  to  suit  either  direction  of  rotation, 
according  to  the  purchaser's  wishes.  The  chief  features  which  render  the  una- 
flow  engine  of  particular  value  for  locomobiles  are  the  central  mounting  of  the 
cylinder  on  the  boiler,  the  arrangement  of  the  valve  gear  on  the  cylinder,  the 
well  balanced  connection  between  the  shaft  bearings  through  the  forked  frame 
to  the  cylinder,  the  freedom  of  relative  expansion  between  boiler  and  frame,  and 
the  simplicity  of  the  entire  construction. 

The  excellent  superheater  designs  of  the  Maschinenfabrik  Badenia  shown  in 
Figs.  9  and  10  are  worthy  of  note.  The  advantages  of  this  design  are  small  thrott- 
ling losses,  large  capacity  per  unit  of  surface,  and  accessibility  of  the  superheater 
and  boiler  tubes.  The  superheater  is  mounted  at  the  smoke  box  end  of  the  boiler 
in  such  a  manner  that  the  area  in  front  of  the  flue  tubes  is  left  free  for  access  to 
the  latter  (Fig.  9). 


256 


Fig.  5. 


Fig.  6. 


Fig.  7. 


257 


Stump/,  The  una-flow  steam  engine. 


17 


258 


Fig.  9. 


Fig.  10. 


259 

In  the  construction  shown  in  Fig.  10  the  same  result  is  attained  by  arranging 
the  superheater  tubes  in  a  winding  fashion  back  and  forth  between  the  flues,  so 
that  the  latter  are  still  accessible. 

Figs.  11,  12  and  13  illustrate  later  designs  by  the  same  builders  which  show 
greater  care  in  the  support  of  the  crosshead  guides  and  the  main  bearings,  as  weli 


bD 


as  increased  area  of  the  exhaust  openings.  The  unbalanced  vertical  centrifugal 
forces  of  the  counterweights  are  transmitted  from  the  main  bearings  through 
supporting  colums  to  the  foundations.  More  latitude  is  therefore  allowed  in  the 
proportioning  of  the  counterweights. 

19* 


260 


Fig.  12. 


Fig.  13. 


261 


Fig.  14. 


Fig.  15. 


262 

In  Fig.  14  is  shown  a  una-flow  semi-portable  engine  or  locomobile,  built  by 
Robey  &  Co.,  Ltd.,  of  Lincoln,  England. 

Figs.  15  and  16  illustrate  a  una-flow  locomobile  constructed  by  the  Erste 
Briinner  Maschinenfabrik,  of  Briinn,  Czecho- Slovakia. 

In  Figs.  17  to  19  are  shown  the  general  arrangement,  cylinder  design  and  valve 
gear  diagrams  of  an  agricultural  una-flow  portable  engine  built  by  the  Kolomna 


Fig.  16. 


Engine  Works,  of  Kolomna,  near  Moscow.  This  type  of  engine  must  necessarily 
be  of  cheap  and  simple  construction,  and  yet  answer  all  the  demands  put  upon  it. 
With  this  in  view,  the  single-beat  valves  are  arranged  horizontally,  with  an  ope- 
rating mechanism  common  to  both.  This  consists  of  a  three-armed  rocking  cam 
lever,  one  arm  of  which  projects  between  the  ends  of  the  valve  stems.  The  others 
are  formed  with  cam  profiles,  by  means  of  which  it  is  rocked  back  and  forth  by 
contact  with  the  roller  of  a  swinging  lever  mounted  on  a  shaft  above  it.  The 


263 

latter  is  operated  by  a  lever  and  link  from  a  shifting  eccentric  controlled  by  a  shaft 
governor.  The  whole  of  the  cam  mechanism  runs  in  oil  and  is  enclosed  in  a  housing 
cast  onto  the  exhaust  belt.  The  cover  of  the  housing  is  adapted  to  support  the 
smoke  stack  of  the  boiler  when  it  is  not  in  use.  In  this  small  type  of  engine,  un- 
balanced single-beat  valves  are  used,  and  these  are  made  in  one  piece  with  their 


rn 


stems.  The  horizontal  arrangement  of  the  valves  was  adopted  in  order  to  obtain 
a  simple  drive  common  to  both  of  them.  The  cylinder  is  cast  in  one  piece  with  the 
crank  end  head. 

This  engine  was  designed  for  use  with  saturated  steam  of  10  at.  pressure. 
In  accordance  with  the  results  of  previous  tests,  the  heads  as  well  as  the  ends 
of  the  cylinder  barrel  were  well  jacketed.  A  neutral,  unheated  zone  extends 
between  the  cylinder  jackets  and  the  central  exhaust  belt. 


264 


The  cylinder  rests  on  a  casting  (Fig.  20)  which  at  the  same  time  forms  a 
housing  for  the  stop  valve  and  distributes  the  steam  to  the  ends  of  the  cylinder. 
The  passages  in  this  casting  are  so  arranged  that  the  water  of  condensation  from 

the  jackets  may  run  back  to  the 
boiler. 

The  piston  is  made  in  two 
parts,  with  bowl  -  shaped  ends 
to  accommodate  the  necessary 
clearance  space. 

In  Figs.  21  to  26  are  shown 
details  of  the  crosshead  guides, 
stuffing  box,  crosshead,  connec- 
ting rod,  main  bearings  and 
crank  shaft. 

The  boiler  is  provided  with 
a  fire  box  of  large  size,  so  that 
straw  and  wood  may  be  used 
for  fuel. 

The  assembly  drawing  of 
Fig.  17  and  the  half-tone  illu- 
stration of  Fig.  27  show  that 
the  construction  is  extremely 
simple  and  well  adapted  to  ser- 
vice conditions.  All  the  previous 
experience  and  results  of  tests 
have  been  fully  utilized  in  order 
to  obtain  a  very  low  steam 
consumption. 

A  similar  portable  engine 
was  built  by  the  Engine  Works 
of  the  Hungarian  State  Rail- 
ways in  Budapest.  A  pair  of  in- 
dicator cards  of  this  engine  are 
reproduced  in  Fig.  28,  and  the 
cylinder  is  shown  in  Fig.  30.  To 
permit  of  the  use  of  a  small 
clearance  volume,  steam  operated 
auxiliary  exhaust  valves  (Fig.  29) 
are  provided  near  the  ends  of 
the  piston  travel.  To  the  end  of 
each  valve  spindle  is  attached  a 

spring  loaded  piston,  the  outer  side  of  which  is  connected  by  a  pipe  to 
the  clearance  space  of  the  corresponding  cylinder  end.  During  admission  and 
expansion,  the  high  steam  pressure  acting  on  the  valve  piston  holds  the  valve 
closed,  but  as  soon  as  the  pressure  is  relieved  through  the  exhaust  ports,  the  valve 


265 

is  opened  by  its  spring.  The  steam  remaining  in  the  cylinder  is  then  swept  out 
by  the  piston  until  the  latter  overruns  the  passage  leading  to  the  auxiliary  exhaust 
valve,  and  compression  begins.  The  compression  pressure,  assisted  by  the  sub- 
sequent admission  of  live  steam,  then  closes  the  valve.  Steam  dashpots  are  used  to 
make  these  valves  quiet  in  operation.  The  piston  rod  stuffing  box  is  also  worthy 
of  notice.  All  the  details  of  this  engine,  including  the  auxiliary  exhaust  valves, 
have  proved  very  satisfactory  in  service.  '  The  action  of  the  exhaust  valves,  as 


HSv.H 


Mean  position  of  eccentric  rod 


Fig.  19. 


well  as  the  effect  of  too  large  an  exhaust  port  area,  is  distinctly  noticeable  in  the 
indicator  cards  of  Fig.  28.  The  steam  consumption  is  low,  as  is  shown  by  the  follow- 
ing test  results: 

Date Feb.  25,  1914 

Duration  of  test .  min     420 

Feed  water  used,  total kg  1424 

„      per  hour „      203.42 

,,          ,,          ,,        ,,       ,,  and  per  1  sqm  heating  surface     „        24.1 

Temperature  in  feed  water  tank,  mean °C       25.5 

Goal  used,  total kg    400 

,,         ,,      per  hour „        57.14 

,,         ,.        ,,       ,,      &  per  1  sqm  grate  area,  mean .    .      ,,      168 


266 


Fig.  20. 


Fig.  21. 


267 


Fig.  2' 


Fig.  25. 


Fig.  26. 


268 


Fig  27. 

Ash,  clinkers,  etc.,  total kg  46.5 

„      %  of  coal  used %  11.6 

Steam  pressure,  kg/sqcm  gage     .    , 10 

Temperature  of  air  at  fire  door      °C  22 

,,             ,,    gases  in  smoke  box,  mean ,,  540 

Draft  in  smoke  box,  mean mm  of  water  10.2 

Evaporation  per  kg  coal kg  3.56 

Steam  consumption,  total ,,  1424 

„                  ,,             per  hour,  mean ,,  203.42 

Revolutions  per  minute 249.5 

Brake  load kg  93 

Lever  arm  of  brake  load mm  560 

Indicated  HP 21.60 

Brake  HP 18.13 

Mechanical  efficiency 0.84 

Coal      consumption  per  B  HP-hour kg  3.15 

Steam            „              „     BHP-   „        „  11.22 

Steam            „              „     JHP-    „        „  9.40 

The  more  than  ample  exhaust  port  area  of  this  and  of  other  cylinders  designed 
by  the  author  finally  led  him  to  the  calculation  of  the  exhaust  and  inlet  areas, 
which  was  given  in  chapter  I,  3  a,  in  dealing  with  throttling  losses.  The  great 


269 

influence  of  the  back  pressure  of  the  exhaust  and  of  the  lead  on  the  port  area 
was  thus  recognized.  A  non-condensing  engine  requires  a  much  smaller  port  area 
than  a  condensing  engine.  In  the  design  of  a  non- condensing  cylinder  with  an 


Fig.  28. 


exhaust  lead  of  25  to  30%,  the  exhaust 
belt  shrinks  to  two  narrow  slits  formed 
as  nozzles  as  shown  in  Fig.  31,  to  which 
are  connected  the  two  exhaust  pipes. 
These  appear  surprisingly  small,  but  have 
sufficient  area  (Fig.  32).  The  large  ex- 
haust lead  gives  a  shorter  length  of  com- 
pression, and  therewith  a  smaller  clea- 
rance volume,  so  that  the  auxiliary  exhaust 
valves  of  Fig.  29  and  30  may  be  dis- 
pensed with,  especially  if  the  parts  are  so 
dimensioned  that  a  suction  effect  with 
a  consequent  reduction  of  back  pressure 
is  obtained.  The  proportions  are  chosen 
in  such  a  way  that  the  toe  of  the  diagram 
becomes  rounded  off  instead  of  being 
sharp  like  that  of  Fig.  28.  The  piston 
and  cylinder  thus  become  considerably 
shorter,  and  the  cylinder  is  supported  on 
a  single  foot  only.  The  joint  between 
cylinder  and  cover  is  also  arranged  in  a 
better  manner,  and  the  accessibility  of 
the  valves  is  improved  by  the  use  of  a 
slipper  type  of  crosshead.  A  better  draft 
in  the  firebox  is  obtained  by  the  ex-  Fig.  29. 

pansion    and    diffusion    of    the    exhaust, 

and    finally    all    the    good    constructional    features    of     the    engine    previously 
described    are    retained,    especially    the    valve    gear    and    valve    arrangement. 


270 


Fig.  30. 


Fig.  31. 


271 


Fig.  32. 


272 

Fig.  33  shows  a  four  cylinder  V  type  una-flow  engine  for  automotive  purposes 
designed  by  the  Stumpf  Una-Flow  Engine  Company,  Inc.,  of  Syracuse,  N.  Y., 
three  of  which  have  been  built  in  different  sizes.  The  cylinders  are  cast  in  pairs, 
which  are  arranged  at  90°  with  one  another.  The  two  cranks  are  set  at  180°  and 
two  connecting  rods  work  side  by  side  on  the  same  crank  pin,  the  cylinder  blocks 
being  displaced  axially  for  this  purpose.  The  cylinder  bore  is  85  mm  (33/8")  and 
the  stroke  95  mm  (3%"-).  The  single-beat  valves  are  operated  by  two  cam  shafts 
which  are  movable  endways.  These  cam  shafts  are  provided  with  a  neutral  cam, 
and  cams  for  9%,  25%  and  80%  forward  cut-off,  and  one  for  80%  cut-off  for 
reversing.  All  the  working  parts  are  enclosed. 


273 


V.  The  Una-Flow  Marine  Engine. 

In  recent  marine  practice,  serious  endeavors  have  been  made  to  introduce 
superheating  and  to  employ  balanced  lift  or  poppet  valves  for  steam  distribution. 
The  una-flow  engine  is  especially  adapted  to  meet  this  modern  tendency,  since 
it  is  well  suited  for  such  conditions.  It  is  also  apparent  that  the  advantages  which 
balanced  poppet  valves  have  over  slide  or  piston  valves  are  much  enhanced  when 
exhaust  valves  are  dispensed  with  altogether,  as  is  the  case  with  una-flow  engines. 
At  the  same  time  this  avoids  the  undesirable  complication  of  the  valve  gear, 
which  has  been  the  great  stumbling  block  in  the  introduction  of  balanced  valves 
in  marine  engines.  The  simplification  associated  with  the  una-flow  engine  is  of 
especial  advantage  when  superheated  steam  is  to  be  used,  since  it  is  particularly 
suited  for  this.  Owing  to  the  unequal  distribution  of  superheat  in  the  case  of 
multi-stage  engines,  many  difficulties  have  arisen  in  practice,  chiefly  in  connection 
with  high  pressure  cylinders.  Such  troubles  are  not  likely  to  occur  in  the  una- 
flow  engine,  because  the  superheat  benefits  the  whole  working  cycle,  despite  the 
fact  that  the  latter  extends  into  the  saturated  region.  The  increase  in  reliability 
consequent  thereon  is  of  particular  importance  from  a  marine  standpoint.  In  the 
case  of  the  first  few  una-flow  marine  engines  which  were  constructed,  the  decision 
to  employ  this  type  was  governed  principally  by  the  fact  that  the  problem  of 
introducing  superheated  steam  into  marine  practice  could  be  solved  in  the  simplest 
and  surest  manner  by  its  adoption. 

Since  that  time,  other  experiments  have  shown  that  the  una-flow  engine  is 
also  well  adapted  for  use  with  saturated  steam,  and  it  should  therefore  satisfy  the 
natural  conservatism  of  those  ship  owners  and  engineers  who  still  regard  the 
introduction  of  superheating  with  much  scepticism.  Such  engineers  always  refer 
to  the  absolute  necessity  of  thorough  cylinder  lubrication  which  is  indispensible 
with  high  superheats,  and  which  endangers  the  safe  and  efficient  operation  of 
the  boiler.  This  is  especially  the  case  in  multi-stage  engines,  where  the  most  diffi- 
cult working  conditions  occur  in  the  high  pressure  cylinder,  which  must  be  espe- 
cially well  lubricated.  On  the  other  hand  when  saturated  steam  is  employed, 
cylinder  lubrication  is  commonly  dispensed  with  altogether,  or  else  oil  is  fed  very 
sparingly  and  mostly  at  the  beginning  and  end  of  a  trip.  Such  a  practice  would 
be  better  justified  in  a  una-flow  engine  working  with  saturated  steam,  where 
moreover,  the  balanced  inlet  valves  do  not  require  lubrication. 

The  first  una-flow  marine  engine,  intended  for  a  steam  trawler,  was  built  by 
J.  Frerichs  &  Co.,  A.  G.,  of  Osterholz-Scharmbeck.  This  engine  is  of  450  BHP, 
with  two  cranks  at  90°,  so  that  there  is  a  gain  of  space  for  fish  storage  purposes 
corresponding  to  that  occupied  by  one  of  the  cylinders  of  a  triple  expansion 
engine,  which  had  so  far  been  the  type  usually  employed  for  this  purpose.  Saeuber- 

Slumvf,  The  una-flow  steam  engine.  18 


274 

lich's  patent  valve  gear  was  used,  which  gives  up  to  80%  maximum  cut-off  for 
maneuvering  in  addition  to  ample  valve  lifts  at  normal  cut-offs.  The  engine  works 
with  highly  superheated  steam  and  has  satisfied  every  demand  put  upon  it. 


Fig.  l. 

The  Stettiner  Maschinenbau-A.-G.  Vulkan,  of  Stettin-Bredow,  next  decided 
to  construct  a  una-flow  engine  and  install  it  in  a  steamer  of  their  own  build  (see 
Figs.  2  to  4).  This  engine  has  two  cranks  at  90°,  and  a  cylinder  bore  of  580  mm 
and  a  stroke  of  600  mm.  It  develops  400  BHP  when  using  steam  of  12  at.  gage 
pressure.  The  boilers  are  fitted  with  superheaters,  and  a  mixing  tube  is  provided 
so  that  saturated  steam  may  be  mixed  with  superheated  steam  so  as  to  obtain 


275 

a  fairly  wide  range  of  working  superheats.  A  Klug  type  of  valve  gear  is  employed, 
in  which  the  motion  of  the  end  of  the  eccentric  rod  is  communicated  to  the  hori- 
zontal valves  in  the  cylinder  heads  by  means  of  a  curved  rod  and  a  cam  and  roller 
mechanism.  The  gear  was  designed  for  a  maximum  cut-off  of  only  26%  so  as  to 
obtain  large  opening  of  the  valves  at  early  cut-off.  In  order  to  provide  for  starting 
and  maneuvering,  an  auxiliary  piston  valve  giving  a  maximum  cut-off  of  90% 


Fig.  2. 

is  mounted  on*  the  exhaust  belt,  and  is  operated  from  a  second  pin,  which  in  this 
particular  case  coincides  with  the  point  of  suspension  of  the  arm  of  the  eccentric. 
This  valve  at  the  same  time  controls  a  set  of  auxiliary  jexhaust  ports  to  permit 
of  relieving  the  compression  when  starting  up  with  no  vacuum  in  the  condenser, 
since  the  air  pump  is  directly  driven  from  the  engine.  This  type  of  auxiliary  valve 
gear  is  shown  diagrammatically  in  Fig.  5,  which  represents  the  arrangement  em- 
ployed on  a  una-flow  marine  engine  installed  in  the  steamship  "Strassburg"  owned 
by  the  Hamburg- American  Line.  It  should  be  noted  that  the  pilot  valve  which 
admits  live  steam  to  the  auxiliary  piston  valve,  as  well  as  the  cylinder  valves  which 
control  the  connections  from  the  latter  to  the  ends  of  the  working  cylinder,  are 
actuated  automatically  by  the  valve  gear,  so  that  no  extra  manual  operation  is 
required  for  maneuvering.  When  the  gear  is  in  either  of  its  outermost  positions 

18* 


276 

for  ahead  or  astern  running,  the  pilot  valve  and  cylinder  valves  are  opened,  while 
in  intermediate  positions  of  the  gear  all  these  valves  are  closed.  When  maneuvering, 
it  is  only  necessary  to  turn  the  reversing  wheel  until  the  engine  responds.  If  the 
main  gear  does  not  start  the  engine,  then  the  auxiliary  gear  will  come  into  action. 
As  soon  as  the  engine  begins  to  turn  over,  the  gear  is  brought  back  to  the  normal 
running  cut-off  of  10%.  In  this  position  the  auxiliary  valve  gear  is  completely  cut 
out.  The  elimination  of  hand-operated  valves  thus  greatly  simplifies  maneuvering. 


Fig.  3. 

The  air  pump  and  auxiliaries  are  directly  driven  from  the  main  engine,  and 
it  is  for  this  reason  that  the  auxiliary  piston  valve  is  also  arranged  to  relieve  the 
compression  when  starting  and  maneuvering. 

The  condenser  is  incorporated  in  the  rear  columns  in  the  customary  manner. 
The  entire  valve  gear  is  mounted  at  the  front  of  the  engine  where  all  parts  are  acces- 
sible. As  shown  in  Fig.  6  the  crank  cheeks  are  formed  as  eccentrics.  The  engine 
is  designed  to  work  ordinarily  with  superheated  steam  of  a  temperature  of  only 
250°  G  and  the  ends  of  the  cylinder  barrels  are  therefore  steam  jacketed,  in  addi- 


277 


Fig.  4. 


tion  to  the  heads.  The  cylinders  are  bolted  together  along  their  exhaust  belts, 
where  the  temperature  is  lowest;  the  rigidity  of  the  wrhole  structure  is  therefore 
considerably  increased  without  appreciable  changes  of  alignment  due  to  expansion. 

Fig.  7  shows  the  steam  valve  to- 
gether writh  its  valve  bonnet  and  cam 
mechanism.  In  Fig,  8  is  shown  an  out- 
line view  of  the  cargo  steamer  "Vulkan", 
which  was  fitted  with  the  engine  just 
described. 

The  Hamburg- American  Line  also 
decided  to  fit  two  una-flow  engines  to 
the  twin-screw  steamer  "Strassburg", 
which  plies  between  Hamburg  and  Co- 
logne. This  boat,  shown  in  Fig.  9,  was 
built  in  the  yards  of  Gebriider  Sachsen- 
berg  A.-G.,  of  Deutz  near  Cologne.  Eaeh 
propeller  shaft  is  driven  by  a  two-cylin- 
der vertical  una-flow  engine,  the  cylin- 
ders of  which  have  a  bore  of  440  mm 
and  a  stroke  of  450  mm.  Each  engine 
develops  250  I  HP  at  175  r.  p.  m.  when 
using  steam  of  12  at.  pressure,  at  a  tem- 
perature of  325°  C.  As  will  be  seen  from 
the  photograph  reproduced  in  Fig.  10, 
and  from  the  drawings  given  in  Fig.  11, 
these  engines  are  built  on  the  same  lines 
as  the  "Vulkan"  engine  just  described. 
They  are  likewise  fitted  with  the  Klug 
type  of  valve  gear  with  auxiliary  gear 
as  described  above.  On  account  of  the 
high  degree  of  superheat,  the  steam 
jackets  on  the  ends  of  the  cylinder 
barrel  were  omitted. 

The  firm  of  Burmeister  &  Wain,  of 
Copenhagen,  also  decided  to  introduce 
the  una-flow  engine  on  two  single-screw 
steamers  ordered  by  the  United  Steam- 
ship Co.,  of  Copenhagen.  Each  engine  is  of  1000  BHP  and  has  three 
cylinders.  The  Klug  type  of  valve  gear  is  employed,  but  the  auxiliary 
mechanism  is  omitted  since  with  three  cranks  at  120°  the  maximum  cut-off 
necessary  for  maneuvering  is  only  about  40%.  With  this  longest  cut-off 
the  inlet  valves  still  give  sufficient  opening  at  the  normal  cut-off  of  10%. 
Each  cylinder  has  independent  steam  and  condenser  connections,  and  it 
thus  becomes  possible  to  cut  out  any  one  in  case  of  need.  By  using  a  longer 
cut-off  in  the  two  remaining  cylinders,  the  full  running  power  may  then  be 
obtained. 


Fig.  5. 


279 


Fig.  6. 


Fig.  7- 


280 

Half  side  views  of  these  engines  are  shown  in  Figs.  12  and  13.  These  photo- 
graphs clearly  show  the  compact  arrangement  of  the  Klug  gear  at  the  front  end 
of  the  engines,  the  three  operating  rods  being  arranged  to  rock  three  telescopic 
shafts  which  transmit  the  motion  to  the  valves  of  the  individual  cylinders.  The 


Fig.  8. 


Fig.  9. 

latter  have  a  bore  of  635  mm,  with  a  stroke  of  915  mm,  the  speed  being  84  r  p.  m. 
The  cylinders  are  designed  to  give  a  very  compact  assembly,  so  that  a  total  saving 
of  1.75  m  in  the  length  of  the  engine  results.  This  arrangement  has  the  advantage 
of  giving  a  better  static  balance  of  the  reciprocating  parts,  so  that  the  cut-offs 
at  both  head  and  crank  ends  of  each  cylinder  may  be  made  equal.  This  results 
in  very  smooth  running.  The  consideration  of  dynamic  balance  is  unimportaut 
at  the  low  speed  in  question.  The  air  pump  is  directly  driven  from  the  main 
engine,  and  a  small  auxiliary  pump  is  provided  for  creating  a  vacuum  before  starting 


281 

the  engine.   This  may  also  be  accomplished  by  using  an  ejector  or  by  opening  the 
blow-off  cocks. 

All  the  cylinders   are  provided  with  bleeder  valves,  the  steam  withdrawn 
being  used  for  heating  the  feed  water. 


Fig.  10. 

The  connection  between  the  exhaust  belts  of  the  cylinders  and  the  condenser 
is  especially  worthy  of  notice.  Each  cylinder  has  an  individual  connection,  but 
since  the  exhaust  belts  are  all  interconnected,  there  is  always  ample  area  of  pas- 
sage from  each  cylinder  to  the  condenser. 


282 

Despite  the  high  superheat  used,  there  is  no  special  provision  for  cylinder 
lubrication.  This  is  due  to  the  fact  that  with  single-stage  expansion,  even  with 
high  superheats,  the  cycle  extends  into  the  saturated  region,  so  that  the  average 
temperature  of  the  working  surfaces  is  low.  The  cylinders  are  lubricated  sparingly 
only  at  the  beginning  and  end  of  a  trip,  while  in  ordinary  running  they  are  not 


oiled  at  all.  For  this  reason  it 
was  not  thought  necessary  to 
provide  an  oil  separator. 

One  test,  which  was  not 
merely  an  exhibition,  but  a  test 
under  actual  working  conditions, 
showed  a  coal  consumption  of 
0.6  kg/I  HP-hour.  The  coal  used 
was  Newcastle  coal  having  a 
calorific  value  of  7300  cal/kg. 
The  steam  pressure  was  11  at. 
and  the  temperature  220°  C.  The 
steam  used  by  the  auxiliary 
machinery  was  included  in  figuring  the  above  result.  It  should  be  noted  that  no 
forced  draft  or  means  of  preheating  the  air  is  provided,  and  that  the  ends  of  the 
cylinder  barrels  are  not  jacketed.  The  boilers  are  fitted  with  Jorgensen  patent 
superheaters  which  have  proved  very  reliable  in  service. 

The  steamer  proved  most  satisfactory  to  both  the  purchasers  and  builders, 
and  it  was  a  pleasant  surprise  to  both  parties  to  find  that  the  guaranteed  speed 


Fig.  11. 


283 


Fig.  12. 


284 


Fig.  13. 


285 

was  obtained  with  800  HP  instead  of  1000  HP  as  mentioned  in  the  specification. 
The  excellent  indicator  cards  taken  during  the  trial  trip  are  shown  in  Fig.  14. 

Fig.  15  shows  a  design  of  a  una-flow  marine  engine  embodying  a  Walschaert 
valve  gear,  which  has  been  so  successful  on  locomotives. 

In  Fig.  16  is  shown  the  assembly  of  the  compound  engine  of  the  steamer 
"Wera",  owned  by  the  Orient  Co.,  of  Petrograd.  On  account  of  the  excellent 
steam  consumption  results  obtainable  with  the  una-flow  engine,  this  company 
decided  to  replace  the  compound  engine  by  a  two  cylinder  una-flow  machine 
having  a  cylinder  bore  of  600  mm  and  a  stroke  of  711  mm.  Superheating  had 
been  tried  experimentally  on  this  ship,  but  the  old  engine  had  proved  unsuitable 
for  use  with  it.  For  this  reason  it  was  decided  to  change  over  to  una-flow  cylin- 
ders, the  design  being  shown  in  Figs.  17  and  18.  The  steamer  is  fitted  with  two 
engines,  each  of  which  is  capable  of  developing  500  HP  at  a  maximum  speed  of 
125  r.  p.  m.  The  ends  of  the  cylinder  barrels  are  steam  jacketed,  so  that  the  engine 
is  suitable  for  working  with  saturated  steam  if  the  occasion  should  arise. 

In  marine  practice  there  is  also  need  for  a  valve  gear  which  will  give  proper 
valve  lifts  for  normal  cut-offs  without  having  excessive  movement  for  cut-offs 
of  70  to  80%.  From  this  point  of  view  was  developed  the  valve  gear  shown  in 
Fig.  19. 

In  all  the  common  valve  and  reversing  gears  the  angle  of  advance  as  well 
as  the  throw  of  the  resultant  eccentric  are  varied,  with  the  effect  that  for  small 
cut-offs  the  throw  of  the  eccentric  is  also  small,  as  is  the  valve  opening.  All  parts 
of  the  engine,  however,  have  to  be  proportioned  to  accommodate  the  maximum 
eccentric  throw.  Obviously,  the  valve  opening  for  small  cut-offs  could  be  mate- 
rially improved  if  instead  of  changing  the  eccentricity  and  angle  of  advance,  the 
former  is  kept  constant  and  only  the  latter  is  changed.  This  method  of  course 
necessitates  a  change  of  the  lap  lines  as  is  shown  in  Fig.  22,  in  which  the  left  hand 
diagram  is  drawn  for  valve  gears  of  the  Klug  or  Walschaert  type  while  the  right 
hand  diagram  is  for  a  gear  incorporating  this  new  principle.  The  right  hand  dia- 
gram is  drawn  so  that  the  constant  throw  of  the  eccentric  is  equal  to  the  maxi- 
mum throw  in  the  case  of  the  left  hand  diagram.  The  valve  opening  "A"  for 
maximum  cut-off  is  equal  in  the  two  cases.  For  normal  cut-off,  however,  the 
valve  opening  "B"  of  the  new  type  of  valve  gear  is  seen  to  be  considerably  larger 
than  that  given  by  the  standard  valve  gears. 

In  the  design  shown  in  Fig.  19,  a  bevel  gear  is  keyed  to  the  end  of  the  crank 
shaft  and  meshes  with  a  pinion  mounted  in  a  rotatable  housing,  which  in  turn 
drives  another  bevel  gear  similar  to  the  first,  but  in  the  opposite  direction.  The 
latter  is  keyed  to  a  sleeve  together  with  four  eccentrics,  each  of  which  operates 
one  of  the  valves  of  the  two  cylinders.  These  eccentrics  are  connected  to  rocking 
levers  mounted  on  an  eccentric  spindle  which  is  geared  to,  and  is  turned  simul- 
taneously with  the  gear  housing.  This  gear  housing  is  turned  by  means  of  the 
reversing  wheel,  and  in  order  to  turn  the  eccentrics  and  the  eccentric  rocker  spindle 
through  the  same  angle,  the  gear  ratio,  on  account  of  the  differential  action,  must 
be  two  to  one.  The  main  eccentrics  as  well  as  the  eccentric  rocker  spindle  are 
moved  in  the  same  direction  and  in  such  a  relationship  that  with  one  of  the  cranks 
in  its  dead  center  the  corresponding  cam  roller  remains  stationary,  thereby  keeping 


286 


288 


o 

^* 

si 


289 


the  point  of  opening  constant.  When  the  main  eccentrics  are  thrown  from  full 
ahead  to  full  astern,  the  eccentric  rocker  spindle  is  moved  through  the  same  angle. 
The  resultant  eccentric  curve  for  this  gear  is  therefore  not  a  straight  line  but  a 
circle  drawn  about  the  center  of  the  shaft.  The  throw  of  the  eccentric  is  always 
the  same  and  the  lap  is  altered  in  proportion  to  the  angle  of  advance.  When  the 
arms  of  the  rocking  levers  have  the  ratio  1:1,  the  eccentricity  of  the  spindle 
must  be  one-half  that  of  the  main  eccentrics.  The  eccentric  rods  are  very  short 
in  order  to  compensate  for  the  angularity  of  the  connecting  rods.  In  addition, 
the  rocking  levers  are  proportioned  to  give  a  later  cut-off  at  the  crank  ends  of  the 
cylinders  and  an  earlier  cut-off  at 
the  head  ends  for  forward  running, 
so  as  to  eliminate  the  effect  of  the 
weight  of  the  reciprocating  parts. 

The  motion  of  the  valves  is 
derived  from  that  of  the  rocking 
levers  by  means  of  reach  rods  and 
cam  and  roller  mechanisms.  The 
rollers  of  the  head  and  crank  end 
bonnets  are  arranged  above  the 
cams,  and  the  eccentrics  are  set  so 
that  the  rocking  levers  for  the  same 
cylinder  move  in  opposite  directions. 

An  interesting  modification  of 
this  gear  is  shown  in  Fig.  20.  In 
this  case  the  eccentrics  operate  the 
four  valves  of  the  two  cylinders 
through  telescopic  shafts.  The  latter 
are  mounted  on  a  long  spindle 
mounted  eccentrically  in  two  bea- 
rings arranged  on  the  exhaust  belt. 
This  spindle  and  the  bevel  gear 
housing  are  interconnected  by  a 
vertical  spindle  and  two  worm 
gears  in  such  a  manner  that  the 
center  of  the  telescopic  shafts  is  swung  through  the  same  angle  as  the  main  eccen- 
trics. When  the  latter  are  moved  from  the  full  ahead  to  the  full  astern  position, 
the  center  of  the  telescopic  shafts  is  displaced  by  the  same  angle.  The  throw  of 
the  main  eccentrics  and  the  eccentricity  of  the  spindle  carrying  the  telescopic  shafts 
are  so  proportioned  that  the  lap  of  the  inlet  valves  is  changed  in  conformity  with 
the  alteration  in  the  angle  of  advance  of  the  eccentrics.  In  this  case,  as  in  the 
other  form  of  this  valve  gear  described  above,  no  auxiliary  gear  is  required,  since 
cut-offs  up  to  85%  are  easily  obtainable.  A  diagrammatic  outline  of  this  gear  is 
shown  in  Fig.  21. 

In  Fig.  23  are  shown  a  side  elevation  and  plan  of  a  una-flow  marine  engine 
for  a  paddle  steamer  plying  on  the  river  Volga  in  Russia.  This  engine  has  a  cylinder 
bore  of  600  mm,  a  stroke  of  800  mm  and  develops  180  BHP  at  26  r.  p.  m. 

Stumpf,  The  una-flow  steam  engine.  19 


Fig.  17. 


290 

Details  of  the  cylinders  and  heads  are  given  in  Figs.  24  and  25. 

The  crank  end  cylinder  heads  are  tied  to  the  main  bearings  by  cast  steel  rods 
of  square  section  which  also  serve  as  crosshead  guides.  The  main  bearing  housings 
are  cast  in  one  piece  with  their  supports.  The  valve  gear  is  arranged  at  one  side 
of  the  engine  and  consists  of  two  sets  of  Klug  type  reversing  motions  which  ope- 
rate the  valve  cam  mechanisms  through  the  medium  of  a  pair  of  telescopic  shafts 
mounted  transversely  on  the  exhaust  belts.  The  arms  of  the  eccentric  straps, 
extend  downwards  and  are  connected  at  intermediate  points  by  links  to  a  yoke 
piece  which  may  be  adjusted  by  hand  through  screw  gearing.  The  main  valve 
gears  are  designed  for  a  maximum  cut-off  of  25%,  the  head  and  crank  end 


Fig.  18. 


cut-offs  being  equalized  as  far  as  possible.  An  auxiliary  valve  gear  is  driven 
from  a  pin  on  each  of  the  short  eccentric  arms,  which  in  this  case  coincides 
with  the  point  of  attachment  of  the  swinging  link.  This  auxiliary  gear  gives 
a  cut-off  up  to  90%  of  the  stroke,  and  thereby  permits  of  easy  maneuvering. 
The  auxiliary  valves  are  operated  by  means  of  a  cross  shaft  and  rocking 
levers  supported  at  the  crank  end  of  the  cylinders.  The  auxiliary  gear  is 
cut  out  by  a  suitable  mechanism  actuated  by  a  sleeve  mounted  on  the  cross 
shaft,  this  mechanism  being  connected  to  the  main  valve  gear  yoke  in  such 
a  manner  that  the  entire  control  during  maneuvering  is  effected  from  the 
main  gear. 


291 


o> 

*-H 

tc 


19* 


292 

In  Fig.  26  is  shown  a  modified  design  of  this  engine  incorporating  a  gear 
similar  to  that  shown  in  Fig.  20.  The  telescopic  shafts  are  supported  on  a  long 
eccentric  spindle  arranged  on  the  exhaust  belts  on  top  of  the  cylinders. 


o 
«M 

t'c 


In  Fig.  27  is  illustrated  a  two-cylinder  una-flow  marine  engine  built  by  the 
Kingsford  Foundry  &  Machine  Works,  of  Oswego,  N.  Y.    This  engine  is  fitted  to 


293 


a  tug  boat  working  on  the  New  York  State  barge 
canal,  and  has  a  cylinder  bore  of  18"  with  a 
stroke  of  18".  The  valve  gear  is  arranged  at  the 
front  of  the  engine  and  consists  of  a  revolving 
cam  shaft  driven  from  the  crank  shaft  by  two 
sets  of  spiral  gears  and  an  intermediate  vertical 
shaft.  The  cam  shaft  is  enclosed  in  a  housing 
bolted  to  the  exhaust  belt,  and  operates  the  ver- 
tical valves  in  the  cylinder  heads  by  means  of 
roller  levers  and  tappets.  A  set  of  five  cams  is 
provided  for  each  cylinder,  which  give  cut-offs 
of  75%  astern,  neutral,  75%  ahead  for  maneu- 
vering, full  load  ahead,  and  half  load  ahead. 
The  cam  shaft  is  shifted  bodily  endways  by  le- 
vers actuated  by  a  steam  cylinder,  the  valve  of 
which  is  controlled  by  a  hand  lever  having  a 
follow-up  motion.  Equal  cut-offs  at  the  head 
and  crank  ends  of  the  cylinders  are  obtained  by 
placing  the  roller  for  the  head  end  valve  4%° 
ahead  of  its  normal  180°  position,  the  average 
difference  in  crank  angle  for  equal  cut-offs  being  9°. 
The  tappet  of  the  head  end  valve  has  a  clearance 
sufficient  to  make  4%°  of  the  cam  inoperative, 
so  that  the  total  time  of  opening  of  the  head 
end  valve  corresponds  to  the  cam  angle  minus 
twice  the  angle  of  offset,  or  9°.  The  cams  as 
well  as  the  rollers  are  beveled  off  to  facilitate 
the  endwise  movement  of  the  cam  shaft.  This  en- 
gine has  proved  thoroughly  reliable  in  operation. 


Fig.  21. 


Fig.  22. 


294 


In  Figs.  28,  29  and  30  is  shown  a  four  cylinder  single-acting  una-flow  marine 
engine  having  the  condenser  incorporated  in  the  frame.  This  engine  was  built  by 
the  firm  of  Karl  Schmid  of  Landsberg  for  the  steamer  "Koriolan",  and  has  a 
cylinder  bore  of  470  mm,  a  stroke  of  350  mm,  and  develops  400  HP  at  250  r.  p.  m. 
For  better  balancing,  the  cranks  of  each  pair  of  adjacent  cylinders  are  set  at  180°, 
and  the  pairs  of  cranks  are  in  turn  set  at  90°.  A  separate  crosshead  guide  is  pro- 
vided, apart  from  the  cylinder  bore,  and  the  piston  is  accordingly  of  stepped  con- 
struction. Any  water  of  condensation  dripping  from  the  pistons  is  thus  kept  away 
from  the  lubricating  oil  in  the  crank  pit.  The  cranks  have  scoops  formed  upon 
them  which,  dip  into  the  oil  and  deliver  it  to  the  crank  pins.  A  further  advantage 
claimed  by  the  builders  for  this  type  of  construction  is  the  lower  working  tempe- 
rature of  the  crosshead  piston  and  its  removal  from  the  hot  cylinder  walls.  This 
type  has  proved  very  reliable,  but  the  same  may  be  said  of  the  straight  piston 
design  shown  in  Fig.  31,  if  a  different  method  of  lubrication  is  employed.  The 


Fig.  24. 


Fig.  25. 


valves  are  located  centrally  in  the  heads  and  are  operated  by  cams  through  bell 
crank  levers  mounted  on  eccentric  pivots  by  means  of  which  the  cut-off  may  be 
changed  or  the  valves  made  inoperative  (see  Fig.  30).  Corresponding  to  the  posi- 
tions of  the  hand  lever,  the  cam  has  separate  steps  for  60%  cut-off  astern,  neutral, 
60%,  20%,  10%  and  5%  cut-off  ahead.  The  cam  shaft  runs  at  half  engine  speed, 
for  which  reason  it  is  provided  with  two  sets  of  cam  profiles  for  each  cut-off.  By 
means  of  the  eccentric  adjustment  the  rollers  may  be  moved  horizontally  and 
be  brought  into  proper  engagement  with  the  cams  or  swung  clear  of  them.  The 
cam  shaft  is  shifted  by  means  of  a  handwheel  and  rack  and  pinion.  Each  cylinder 
is  provided  with  a  separate  stop  valve. 

The  general  construction  of  the  engine  and  some  of  the  details  of  the  valve 
gear  recall  those  of  a  marine  oil  engine.  The  surface  condenser  is  incorporated  in 
the  frame,  and  this  arrangement  results  in  a  considerable  saving  of  space  and 
a  reduction  of  the  back  pressure,  although  it  is  probably  only  suitable  for  small 
engines.  This  machine  has  given  excellent  satisfaction. 

The  four  cylinder  marine  engine  shown  in  Fig.  31  is  of  similar  design,  but 
in  this  case  straight  pistons  are  used.  The  valve  gear  shaft  is  also  driven  by  spiral 


Fig.  23. 


295 


Fig.  27. 

gearing,  but  in  contrast  to  that  of  the  engine  just  described,  it  runs  at  double  the 
speed  of  the  crank  shaft.  Single-beat  high  lift  valves  are  arranged  centrally  in 
the  cylinder  heads  and  are  operated  by  Lentz  cam  mechanisms  by  means  of  rocking 
levers  mounted  on  eccentric  pins,  these  levers  being  actuated  by  eccentrics  on 


296 

the  main  valve  gear  shaft.  The  latter  is  driven  through  a  differential  gear,  the 
housing  of  which  may  be  turned  by  a  worm  and  hand  wheel  for  altering  the  phase 
relation  to  the  shaft,  whereby  the  angle  of  advance  and  consequently  the  cut-off 
is  changed  and  the  engine  may  be  reversed.  The  necessary  lap  of  the  valve  cams 


Fig.  28. 


is  obtained  by  communicating  the  rotative  motion  of  the  differential  gear  housing 
to  the  eccentrics  on  which  the  rocking  levers  are  mounted,  in  such  a  way  that 
the  angle  of  advance  of  the  latter  always  corresponds  to  that  of  the  main  eccen- 
trics. The  whole  of  the  valve  gear  is  enclosed  in  a  housing.  In  consequence  of  the 
double  speed  of  the  valve  gear  shaft,  the  angle  of  cut-off  is  doubled  and  the  valve 


297 

lift  quadrupled,  which  thus  permits  of  the  use  of  a  very  small  single-beat  valve. 
The  latter  is  unaffected  by  pressure  and  temperature  changes  and  will  therefore 
remain  permanently  tight.  The  clearance  space  and  harmful  surfaces  are  also 
materially  reduced.  This  design  has  been  developed  in  connection  with  the  single- 
acting  vertical  two  cylinder  engine  shown  in  Fig.  47  of  Ch.  II,  1,  p.  165,  under 
the  heading  of  stationary  engines.  (See  also  the  following  chapter.) 

In  Fig.  32  is  shown  a  design  of  a  single-acting  vertical  six  cylinder  marine 
engine  developed  by  the  Stumpf  Una-Flow  Engine  Company,  Inc.,  of  Syracuse, 
N.  Y.,  in  which  straight  pistons  are  also  used.  The  double-beat  valves  are  arranged 
horizontally  in  the  heads  and  are  operated  by  bell  crank  levers  actuated  by  tapered 
cams  mounted  on  a  cam  shaft  running  at  engine  speed.  The  valves  are  easily 
removable  from  the  opposite  side  of  the  cylinders,  and  each  cylinder  head  is 
detachable  with  the  valve  in  place  without  disconnecting  the  gear,  so  that  the 
piston  is  easily  accessible.  The  cylinders  are  all  bolted  together  at  their  exhaust 


Fig.  29. 


belts,  where  the  temperature  is  lowest,  and  are  tied  to  the  cast  steel  base  plate 
by  substantial  bolts  with  tubular  distance  pieces.  Cast  steel  side  members  are 
arranged  to  take  the  side  thrust  of  the  piston  and  also  permit  of  the  attachment 
of  side  plates  for  enclosing  the  driving  parts.  The  cam  shaft  with  its  tapered  cams 
may  be  moved  bodily  endways  by  means  of  a  hand  wheel  and  worm  gear,  whereby 
the  cut-off  may  be  changed  and  the  engine  reversed.  The  six  cylinders  have  a  bore 
of  22"  with  a  stroke  of  24"  and  develop  5000  HP  at  250  r.  p.  m.  All  problems 
concerning  manufacture  and  operation  have  been  satisfactorily  solved  in  this  design. 

In  all  single-acting  engines  of  this  kind  the  greatest  care  must  be  exercised 
in  the  design  and  manufacture  of  the  piston  and  rings,  which  must  make  a  va- 
cuum-tight joint  with  the  cylinder.  Air  leakage  past  the  piston  into  the  vacuum 
touches  the  engine  at  a  vulnerable  spot. 

In  the  case  of  high  powered  engines,  if  it  is  desired  to  use  the  Schlick  balan- 
cing arrangement,  this  can  be  carried  out  with  a  four  or  six  crank  una-flow  engine. 
A  four  cylinder  engine  of  this  kind  is  shown  in  Fig.  33.  Hollow  pistons  can  be 


298 

made  very  light,  as  was  mentioned  in  dealing  with  locomotive  details.  Such  light 
pistons  might  be  employed  for  the  outside  units,  while  the  necessary  additional 
weight  can  be  easily  added  in  the  cavities  of  the  pistons  of  the  middle  cylinders 
without  alteration  of  any  other  parts. 

The  effect  of  the  inertia  forces  of  the  reciprocating  masses  upon  the  loads 
on  the  driving  parts  is  more  favorable  in  large  una-flow  engines  than  in  the  modern 
triple  or  quadruple  expansion  engines,  where  the  inertia  and  steam  pressures  are 


Fig.  30. 


additive  in  the  latter  part  of  the  stroke.  The  actual  maximum  stresses  occurring 
in  regular  running  are  much  smaller  in  large  una-flow  engines.  At  the  same  time 
piston  speeds  of  5%  m/sec  may  be  employed. 

Investigation  proves  that  for  cut-offs  later  than  normal,  the  load  distribu- 
tion on  the  driving  parts  of  a  una-flow  engine  is  more  even,  while  for  early  cut-off 
and  at  low  speeds  of  revolution  the  multi-stage  engine  is  better  in  this  respect. 

For  small  and  medium  sizes  and  low  speeds,  the  three  cylinder  arrangement 
has  many  advantages,  such  as  better  load  distribution  on  the  driving  parts,  more 
uniform  torque  and  a  smaller  shaft  diameter. 


299 

With  three,  four,  or  more  cylinders,  the  una-flow  marine  engine  offers  a  great 
reserve  of  power.  Since  each  cylinder  and  set  of  driving  parts  forms  a  complete 
unit  in  itself,  any  one  or  more  of  these  units  may  be  disconnected  if  requiring 
repairs,  while  the  cut-off  in  the  remaining  ones  may  be  increased  to  make  up  for 
the  loss  of  the  one  out  of  action.  For  instance,  it  is  possible  to  cut  out  two  units 
of  the  four  cylinder  engine  just  described,  and  to  increase  the  cut-off  in  the  re- 
maining ones  from  10  to  20%  to  make  up  for  the  difference. 


Fig.  31. 


In  comparison  with  that  of  the  multi-stage  engine,  the  reversing  gear  of  the 
una-flow  marine  engine  is  much  more  simple  and  reliable.  In  the  latter  the  pro- 
cess of  reversing  only  applies  to  the  inlet  valves.  Since  in  this  type  of  engine  there 
are  no  intermediate  receiver  pressures  to  be  taken  into  account  when  reversing, 
and  as  the  compression  is  always  constant,  the  difficulties  caused  by  excessive 
compression  pressures  in  the  first  cylinders  of  multi-stage  engines  are  avoided. 
Reversing  takes  place  much  more  smoothly,  especially  since  the  balanced  inlet 
valves  offer  very  little  resistance.  Consequently  the  valve  gear  parts  are  subject 
to  very  little  wear,  as  is  proved  also  by  experience. 


300 

In  a  quadruple  expansion  engine  the  diagram  factor  of  the  indicator  card 
shown  in  Fig.  34  may  be  taken  as  an  average  of  55  to  60%.    The  remainder  is 


Fig.  32. 

lost  by  throttling  in  the  valves  and  pipes,  and  by  condensation  losses.  On  the 
other  hand,  the  diagram  factor  of  the  indicator  card  (Fig.  35)  of  a  una-flow  engine 
may  reach  80%  with  a  good  vacuum,  i.  e.,  a  difference  of  20  to  25%  in  favor  of 


301 


302 

the  una-flow  engine.  This  explains  in  part  the  essentially  smaller  dimensions  of 
a  una-flow  cylinder  in  comparison  with  the  low  pressure  cylinder  of  a  multi-stage 
engine.  This  is  also  to  some  extent  the  reason  for  the  fact  that  the  steam  con- 
sumption of  a  una-flow  engine  is  not  greater  than  that  of  a  quadruple  expansion 
engine  of  equal  power,  both  for  saturated  and  superheated  steam.  - 


Fig.  34. 


Fig.  35. 

By  distributing  the  steam  flow  to  several  cylinders,  smaller  inlet  valve  dimensions 
are  obtained,  in  contrast  to  the  bulky  valves  necessary  with  multi-stage  engines,  in 
which  the  total  working  steam  has  to  pass  from  one  cylinder  to  the  next  in  series. 

Another  valuable  feature  of  the  una-flow  marine  engine  is  the  small  number 
of  spare  parts  required,  since  each  of  them  may  be  used  with,  any  one  of  the 
cylinder  units. 


303 


VI.  The  Una-Flow  Engine  with  Single-Beat  Valves 
and  Double  or  Triple-Speed  Lay  Shaft. 

The  usual  types  of  valve  gears  with  fixed  lap  necessarily  give  very  small  valve 
lifts  at  early  cut-offs.  P'or  instance,  at  10%  cut-off  the  valve  opening  a'  is  only 
0.065  r  (see  Fig.  1).  In  order  to  obtain  the  necessary  valve  area  when  slide  valve 
gears  are  used,  a  large  travel  and  considerable  lap  must  be  provided  and  this 
results  in  large  friction  losses  and  leakage.  In  poppet  valve  gears  very  steep  cams 
become  necessary.  In  order  to  alter  the  cut-off,  the  resultant  eccentricity  has 
to  be  changed;  and  since  cut-offs  up  to  50%  must  be  provided  for  in  many  cases, 
the  eccentric  travel  for  early  cut-offs  is  short  and  therefore  the  resultant  valve 
lift  is  small,  i.  e.,  the 
inadequate  effect  of  the 
above  measures  is  partly 
or  wholly  nullified  by  the 
necessity  for  a  long  range 
of  cut-off. 

A  solution  of  the  pro- 
blem on  this  basis  is  im- 
possible, since  the  power 
necessary  to  lift  the  valve 
through  the  required 
height  in  a  given  time 
cannot  be  applied  to  it 
in  this  way.  Instead  of  Fig.  1. 

trying  to  accomplish  the 

work  with  small  valve  lifts  and  large  forces,  it  would  be  better  to  use  large  valve 
lifts  and  small  forces.  If  therefore  the  lay  shaft  is  arranged  to  run  at  double 
the  engine  speed,  then  the  period  of  valve  opening  extends  through  twice  the 
angle  a  arid  the  opening  increases  from  a'  to  a,  as  shown  in  Fig.  1.  Since 


a  =  2r 


sin2  — ,  therefore  a  =  4  a'  approximately  within  fairly  wide  limits.     By 

£ 


doubling  the  lay  shaft  speed,  the  opening  of  the  valve  is  thus  quadrupled.  In 
order  to  prevent  the  valve  from  opening  twice  during  one  revolution  of  the 
engine,  a  cut-out  eccentric  is  interposed  in  the  valve  gear,  running  at  engine 
speed,  which  increases  the  useful  stroke  of  the  main  eccentric  and  makes  every 
second  stroke  of  the  latter  inoperative.  In  Fig.  2  is  shown  a  Zeuner  diagram  for 
such  a  valve  gear,  in  which  the  two  outstrokes  of  the  resultant  eccentric  are 
shown  in  full  lines  while  the  instrokes  are  dashed.  Every  alternate  outstroke 


304 


23 


produces  a  valve  opening  of  19  mm  while  the  following  stroke  falls  short  of  the 
lap  line  by  4  mm.  The  constructional  simplicity  of  this  gear  is  evident  from  Fig.  3. 
The  double-speed  lay  shaft  carries  a  shaft  governor  which  varies  the  throw 
and  angle  of  advance  of  the  main  eccentric  in  the  usual  manner.  The  latter  ope- 
rates the  cam  lever  in  the  valve  bonnet  indirectly  through  a  double  armed  lever, 

pivoted  on  one  of  the  cut- 
out eccentrics  which  are 
forged  in  one  piece  with 
their  shaft  and  revolve  at 
engine  speed.  The  position 
of  the  gear  shown  in  Fig.  3 
corresponds  to  the  lifting 
stroke,  in  which  the  cut- 
out eccentric  magnifies  the 
motion  of  the  main  eccen- 
tric. After  the  crank  has 
turned  through  180°,  the 
main  eccentric  is  again  in 
the  same  position,  but  the 
cut-out  eccentric  has  also 
turned  through  180°  and 
thus  counteracts  the  motion 
of  the  main  eccentric  with 
the  effect  that  the  valve  re- 
mains  closed.  Figs.  3  and  4 
respectively  show  the  gear  and  the  valve  of  a  una-flow  engine  having  a  cylinder 
bore  of  400  mm,  a  stroke  of  500mm,  and  running  at  150  r.  p.  m.  The  main  di- 
mensions of  the  valve  gear  are  given  in  the  following  table: 


- 

Main 
Eccentric 
Shaft 
(Double  speed) 

Cut-out  Eccentric  Shaft 
Crank  End       Head  End 

Throw  of  eccentric  .    .    .   . 

.    .     mm 

37 

14                    12 

Angle  of  advance      .   .   . 

dpi? 

30 

60                   90 

Lao  . 

mm 

30 

Opening  a  

mm 

20                    17 

Maximum  cut-off  .    .••  .   .    . 

°/n 

21                    24-5 

The  cut-out  eccentric  shaft  is  driven  from  the  lay  shaft  by  a  train  of  spur 
gears;  and  where  auxiliary  exhaust  valves  are  employed  they  may  be  operated 
from  the  former.  Since  the  governor  runs  at  double  the  speed  of  that  of  an  ordi- 
nary lay  shaft  engine,  its  regulating  force  is  quadrupled,  and  it  will  therefore 
hardly  be  affected  by  any  valve  gear  reaction.  The  ratio  between  minimum  and 
maximum  eccentric  travel  of  the  shifting  eccentric  of  an  ordinary  valve  gear  may 
be  designed  to  give  a  maximum  cut-off  of  75%.  If  the  same  ratio  is  used  for  the 
main  eccentric  which  runs  at  double  speed,  the  corresponding  crank  angle  is 


305 

equivalent  to  a  maximum  cut-off  of  only  25%.  This  may  be  improved  upon 
by  a  suitable  change  in  the  angle  of  advance  of  the  cut-out  eccentric,  and 
the  maximum  cut-off  may  thus  be  increased  to  about  35%. 

For  ordinary  stationary  engines  a  maximum  cut-off  of  25%  would  seem  to 
be  sufficient,  since  with  a  normal  cut-off  of  10%  at  rated  load  an  overload  of  more 
than  double  this  amount  may  be  carried.  In  special  cases  the  cut-off  may  be 


Fig.  3. 

materially  increased  by  running  the  main  eccentric  shaft  at  engine  speed  and 
•the  cut-out  eccentric  shaft  at  double  speed  (Fig.  11).  In  this  manner  it  becomes 
possible  to  reach  a  maximum  cut-off  of  70%,  but  in  this  case  the  valve  lift  will 
be  only  doubled  instead  of  quadrupled.  Apart  from  the  increase  in  the  weight 
of  the  valves,  the  forces  necessary  to  accelerate  them  will  consequently  be  doubled. 
Since  these  forces  only  amount  to  a  part  of  the  total  valve  gear  reaction,  the  gover- 
nor, now  revolving  only  at  engine  speed,  should  still  be  able  to  handle  them  satis- 
factorily, provided  that  there  is  sufficient  frictional  resistance  in  the  shifting 
eccentric.  A  design  comprising  a  secondary  eccentric  rotatably  mounted  upon 
a  primary  eccentric  is  suitable  for  this  purpose. 


Stumpf,  The  una-flow  steam  engine. 


20 


306 


The  most  satisfactory  arrangement,  however,  is  to  operate  the  governor  shaft 
at  twice  the  engine  speed,  since  this  quadruples  the  regulating  forces  as  well  as 
the  valve  lift.  The  high  lift  obtainable  in  this  manner  allows  of  the  use  of  a  single- 
beat  valve,  the  diameter  of 
which  need  only  be  one  half 
that  of  the  equivalent  double- 
beat  valve.  Such  a  small 
single-beat  valve  is  extremely 
light  and  permits  of  the  re- 
duction of  the  clearance  vo- 
lume to  a  very  small  amount. 
The  compression  pressure  will 
therefore  run  up  to  a  high  fi- 
gure so  that  the  steam  pres- 
sure on  the  single-beat  valve 
will  be  almost  balanced  at  the 
time  of  opening.  This  con- 
sideration will  show  that  a 
small  clearance  volume  is  ab- 
solutely essential  to  the  use 

of  single-beat  valves  and  that 
Fig.  4. 

the  latter    can    only    be    em- 
ployed   in    combination    with 

the  double-speed  valve  gear.  A  good  vacuum  is  desirable  in  view  of  the  small 
clearance  space.  It  also  demonstrates  why  previous  experiments  with  single-beat 
valves  were  bound  to  fail.  Figs.  4  and  5  show  the  great  simplicity  of  the  single- 
beat  valve  as  compared  with  the  equivalent  double-beat  valve,  both  figures  being 


Fig.  5. 


307 

drawn  to  the  same  scale.  Fig.  6  shows  the  plain  character  of  the  cylinder  head 
castings,  as  well  as  the  short  and  simple  steam  passages  and  the  small  amount 
of  clearance  volume  and  harmful  surfaces.  The  reduction  of  surface  loss,  volume 
loss  and  leakage,  as  well  as  the  losses  due  to  throttling  may  be  expected  to  improve 
the  steam  consumption  by  0.4  to  0.5  kg/HP-hour. 

The  single-beat  valve  has  only  half  the  diameter  of  the  double-beat  valve. 
Despite  the  small  dimensions  of  the  valve,  the  pressure  drop  during  admission 
will  be  less  on  account  of  the  direct  flow  of  steam  with  the  least  possible  changes 
of  area  and  direction,  and  the  well  rounded  corners. 

A  large  nozzle,  in  general,  has  less  friction  losses  than  a  small  nozzle,  since 
the  friction  of  the  walls  is  relatively  less  in  comparison  with  the  quantity  of  steam 
passing  through  it.  For  this  reason  the  friction  losses  of  the  single-beat  valve 
with  its  more  compact  steam  jet  must  be  less  than  those  of  the  double-beat  valve 


Fig.  6. 

where  the  flow  is  split  up.  The  even  profile  of  the  cross-section  of  approach  to 
the  valve  seat  and  the  following  diffusor-like  enlargement  of  the  steam  passage 
will  also  cause  a  gradual  increase  in  kinetic  energy,  with  a  subsequent  change  of 
the  same  into  pressure  energy,  so  that  the  total  amount  of  work  changed  into 
heat  due  to  friction  will  be  small. 

The  reduction  of  throttling  losses  also  benefits  the  governor  action,  since  the 
pressure  difference  at  the  valve  during  the  latter  part  of  admission  constitutes 
the  major  part  of  the  load  to  be  handled  by  the  governor.  In  comparison  with 
this  the  forces  necessary  to  lift  the  valve  are  small,  mainly  because  the  clearance 
space  is  filled  with  steam  at  a  pressure  almost  equal  to  that  of  the  live  steam, 
so  that  an  infinitely  small  lift  of  the  valve  is  sufficient  to  allow  the  pressures  to 
equalize  fully.  On  the  other  hand,  the  increasing  pressure  difference  at  the  valve 
during  closing  imposes  an  increasing  load  upon  the  governor,  if  it  is  not  checked 
by  frictional  resistance  in  the  mechanism.  Since  the  forces  on  the  valve  depend 

90* 


308 

upon  the  lift,  the  diameter,  and  the  pressure  difference,  there  will  be  a  best  dia- 
meter for  which  the  loads  on  the  valve  become  a  minimum. 

A  further  part  of  the  valve  load  is  caused  by  the  valve  spring,  which  should 
therefore  not  be  made  heavier  than  necessary.  The  single-beat  valve,  on  account 
of  its  light  weight,  also  improves  conditions  considerably,  since  the  forces  neces- 
sary to  accelerate  it  are  smaller  than  those  of  the  equivalent  double-beat  valve. 
A  valve  spring  calculation  has  already  been  given  in  Chapter  I,  5,  but  the  method 
there  employed  is  not  applicable  in  this  case,  since  the  combined  motions  of  two 
eccentrics  revolving  at  different  speeds  have  to  be  dealt  with.  A  sketch  of  the 
valve  gear  mechanism  is  shown  in  Fig.  7  and  corresponds  to  the  arrangement 
shown  in  Fig.  3.  Noteworthy  is  the  gentle  rise  of  the  cam  profile  L,  which  in  this 

o  / 

case  has  a  lifting  radius  of  35 ;  -  =  18  mm. 

The  first  step  is  to  draw  up  the  valve  lift  curve.  For  this  purpose  the  move- 
ment of  the  whole  gear  is  determined  point  by  point  for  crank  angles  of  10°  each. 
This  corresponds  to  an  angle  of  20°  at  the  main  eccentric  on  account  of  the  double 
speed  of  the  latter.  The  paths  of  the  points  B  and  D  are  obtained  in  this  manner. 
In  the  present  case  it  will  be  found  that  when  the  crank  has  turned  through  about 
60°,  the  center  K  of  the  roller  is  again  in  the  same  position  as  at  the  start,  and 
the  investigation  will  therefore  be  restricted  to  a  crank  angle  of  from  0°  to  6(R 
The  valve  lift  h  for  any  position  of  the  gear  is  the  distance,  measured  parallel  to 
the  valve  stem  center  line,  between  the  curve  L  and  an  arc  described  with  the  radius 
J K.  The  curve  L  is  drawn  through  the  center  of  the  roller  equidistant  to  the 


309 


310 

cam  profile.    The  valve  lifts  thus  found  are  plotted  in  Fig.  8,  giving  the  curve 
marked  h. 

The  valve  velocities  are  determined  essentially  in  the  same  manner  as  before, 
being  calculated  from  the  angular  velocities  of  the  shafts  and  the  instantaneous 
lever  arms  r1?  r2,  /-3,  r4,  r5.  The  latter  are  obtained  from  the  full  size  drawing. 
The  velocity  of  the  rod  DH  is  combined  of  the  two  velocities  imparted  to  it  by 
the  two  eccentrics.  First  assuming  the  cut-out  shaft  0  to  be  stationary,  then  the 


velocity  of  the  rod  DH  is  v2  =  co 


,  where  co2  is  the  angular  velocity  of  the 


main  eccentric  shaft  and  r1?  r2,  and  r3  are  the  lever  arms.   The  second  component 
of  the  velocity  of  the  rod  DH,  which  is  produced  by  the  cut-out  eccentric,  may 


Fig.  8. 


be  found  by  assuming  the  point  B  to  be  held  stationary.  The  instantaneous  effec- 
tive lever  arm  r6  of  the  cut-out  eccentric  O^C  is  the  distance  of  the  point  0-^  from 
the  line  CG  which  bisects  the  included  angle  between  the  lines  CE  and  CF  drawn 


parallel  to  the  rods  A  B  and  DH.    Very  approximately,  vl  — 


•r. 


It 


should  be  noted  that  r6  changes  its  sign  between  the  crank  angles  0°  and  60°.   The 
real  velocity  of  the  rod  DH  is  v5  =  v±  +  v2.    The  velocity  of  the  valve  is  v  =  v5  •  —• 


311 

In  the  practical  application  of  this  method  it  will  be  found  that  increments 
of  the  crank  angle  of  10°  are  too  coarse  to  permit  of  accurate  determination  of 
the  valve  movement  during  the  lifting  period.  It  is  advisable  to  determine  the 
rapidly  varying  lever  arm  r5  for  every  2  mm  of  roller  travel,  as  is  shown  at  the 
right  in  Fig.  8.  It  will  also  be  found  best  to  continue  the  r5  line  back  to  the  zero 
ordinate  as  shown  dashed.  The  relation  between  the  roller  travel  and  crank  angle 
is  next  determined,  giving  the  curve  a  at  the  right  of  Fig.  8  and  the  calculated 
velocities  v  are  then  plotted  in  the  diagram  at  the  left.  The  points  of  change  of 
curvature  of  curve  r5  and  of  the  velocity  curve  v  in  this  case  correspond  to  a  crank 
angle  of  5°.  The  imaginary  extension  of  the  r5  curve  to  the  zero  ordinate  permits 
of  the  continuation  of  the  velocity  curve  to  zero  crank  angle,  as  indicated  by  a 
dashed  line,  and  thus  facilitates  the  location  of  a  tangent. 


Fig.  9. 

The  tangent  T-T  drawn  at  the  steepest  part  of  the  velocity  curve  gives  the 
maximum  retardation  of  the  valve  to  be  taken  care  of  by  the  spring.  In  the  present 
case,  with  the  engine  running  at  150  r.  p.  m.,  a  crank  angle  of  10°  corresponds  to 

60       10  1 

— : -  •  -r^r  =  —  sec.    From  the  diagram  the  change  of  velocity  for  a  crank  angle 
lou     ot)U        yu 

of  10°  is  found  to  be  0,56  m/sec,   and  therefore  the  acceleration  =  0.56   X  90 
=  50.4  m/sec2.    The  corresponding  force  required  to  accelerate  a  valve  having 

50.4 

a  weight  of  G  kg  is  P  =  G  •  -    -  =  5.14  G. 

9.81 

The  calculation  of  the  valve  spring  is  carried  out  as  shown  in  Chapter  I,  5, 
but  it  should  be  noted  that  for  crank  angles  of  0 — 5°  and  55 — 60°,  the  inertia, 
spring  pressure  and  pressure  on  the  valve  due  to  throttling  act  in  the  same  direc- 
tion, and  are  opposed  only  by  the  steam  pressure  on  the  valve  stem  area.  The 
friction  is  assumed  to  be  balanced  by  the  weight.  The  retardation  during  the 


312 

closing  period  of  the  valve  must  therefore  be  also  calculated  by  means  of 
tangents,  and  in  order  to  keep  it  small,  the  lifting  radius  of  the  cam  should 
be  made  rather  large. 

In  Fig.  9  is  shown  a  una-flow  cylinder  with  auxiliary  exhaust  valves  placed 
near  the  ends  of  the  cylinder  barrel,  both  exhaust  and  steam  valves  being  of  the 


Fig.  10. 

single-beat  type.  The  exhaust  valves  are  opened  after  the  steam  pressure  is  relieved 
by  the  piston  uncovering  the  main  exhaust  ports  and  are  closed  after  the  exhaust 
valve  ports  are  overrun  by  the  piston. 


313 

The  operation  of  the  auxiliary  exhaust  valves  is  thus  simplified,  and  the  single- 
beat  form  becomes  permissible.  They  may  be  operated  by  an  eccentric  placed  90° 
ahead  of,  or  behind  the  main  crank.  In  Fig.  10  the  exhaust  eccentric  is  shown  mounted 
on  the  cut-out  shaft,  and  in  Fig.  1 1  on  the  main  lay  shaft,  both  shafts  running  at  engine 


Fig.  11. 

speed.  The  arrangement  shown  in  Fig.  10,  with  a  double-speed  lay  shaft  and  single- 
speed  cut-out  shaft,  will  give  a  range  of  cut-off  up  to  35%,  while  that  of  Fig.  11, 
with  a  single  speed  lay  shaft  and  double-speed  cut-out  shaft,  has  a  range  up  to  70%. 


314 


Fig.  12. 


315 

• 

An  interesting  design  of  the  self-contained  type  is  shown  in  Fig.  12,  where 
the  crank  shaft  is  used  as  lay  shaft,  the  shifting  eccentric  being  controlled  by 
a  flywheel  governor.  The  auxiliary  eccentric  shaft  is  driven  from  the  crank 
shaft  at  double  engine  speed,  by  spur  gears  placed  at  one  side  of  the  crank, 
while  the  exhaust  eccentric  is  placed  on  the  other  side.  The  movement  of  the 
main  eccentric  is  transmitted  to  both  single-beat  inlet  valves  by  rocking  levers 
as  previously  described,  the  latter  being  pivoted  on  eccentrics  at  180°  on  the 
double-speed  shaft,  thus  magnifying  the  useful  movement.  The  lower  end  of  the 
second  lever  is  moved  by  the  lower  end  of  the  first  one  by  a  pin  with  bushing  and 
sleeve.  The  single-beat  exhaust  valves  are  operated  directly  by  the  exhaust  ec- 
centric. 

The  gear  shown  in  Fig.  11  with  its  long  range  of  cut-off  is  suitable  for  engines 
intended  for  ordinary  driving  purposes,  while  that  shown  in  Fig.  10,  having  a  more 
limited  range  of  cut-off,  is  more  useful  for  those  driving  pumps  and  compressors. 


316 


Summary. 

At  the  beginning  of  this  book  the  different  losses  of  the  steam  engine  were 
analyzed.  The  question  which  arises  at  the  end  is,  how  is  the  engine  to  be  designed 
in  order  to  have  a  minimum  total  of  all  these  losses  ?  Such  an  engine  in  the  form 
of  a  una-flow  engine  with  single-beat  valves  is  presented  in  Fig.  6  of  Ch.  VI,  p.  307. 

The  losses  in  the  steam  engine  are: 

1.  Surface  loss. 

2.  Volume  loss. 

3.  Friction  loss. 

4.  Throttling  loss. 

5.  Leakage  loss. 

6.  Loss  due  to  radiation  and  convection. 

7.  Loss  due  to  incomplete  expansion. 

It  was  demonstrated  that  the  una-flow  engine  with  single-beat  valves  as  shown 
in  Fig.  6  of  Chapter  VI  has  the  smallest  surface  loss.  In  the  first  place  the  extent 
of  the  harmful  surfaces  is  extremely  small.  The  additional  harmful  surface  consi- 
sting of  the  short  and  narrow  steam  inlet  passage  amounts  to  only  about  5%  of 
the  smallest  theoretical  harmful  surface,  i.  e.  twice  the  area  of  the  cylinder  bore. 
In  vertical  engines,  the  additional  harmful  surface  will  be  still  smaller,  as  is  evi- 
dent from  the  two  and  four  cylinder  single-acting  engines  shown  in  Fig.  47, 
Ch.  II,  1,  p.  165,  and  Fig.  31,  Ch.  V,  p.  299.  Furthermore,  the  harmful  surfaces 
may  in  this  case  be  easily  machined  and  thereby  still  further  reduced.  The  part 
of  the  harmful  surface  which  needs  jacketing  the  most,  namely,  the  area  of  the 
cylinder  head,  is  exposed  to  the  heating  action  of  the  highly  superheated  live 
steam  in  the  best  possible  manner.  The  extremely  small  clearance  volume  pro- 
duces a  high  compression  with  a  terminal  compression  temperature  of  about 
900°  C.  The  harmful  surface  is  therefore  thermally  prepared  for  steam  admission 
in  the  best  possible 'manner.  The  surface  loss  must  consequently  be  very  small. 

It  was  also  demonstrated  that  the  una-flow  engine  with  single-beat  valves 
has  the  smallest  volume  loss.  There  is  no  better  way  to  reduce  the  volume  loss 
than  by  keeping  down  the  clearance  volume  to  a  minimum.  The  clearance  volume 
of  the  engine  shown  in  Fig.  6  of  Chapter  VI  is  only  1%,  which  reduces  to  about 
%%  f°r  larger  engines  and  to  about  %%  in  the  case  of  the  vertical  engine  with 
single-beat  valves.  The  long  compression  also  assists  in  further  reducing  the  volume 
loss.  For  these  reasons  the  latter  will  be  extremely  small  despite  the  use  of  single 
stage  expansion.  With  the  small  clearance  volumes  just  given,  and  for  the  usual 
range  of  pressure  drop,  the  critical  back  pressure  of  this  type  of  engine,  i.  e.  the 


317 

back  pressure,  below  which  the  steam  consumption  increases,  is  far  below 
anything  that  can  be  reached  with  even  the  best  condensing  equipment. 

It  was  also  shown  that  the  una-flow  engine  with  single-beat  valves,  assuming 
the  same  lubrication  and  operating  conditions,  has  less  friction  losses  than  the 
equivalent  counterflow  tandem  engine,  since  there  is  only -one  piston  instead  of 
two,  one  piston  rod  packing  instead  of  three,  and  two  valves  instead  of  eight. 

It  was  proved  also  that  the  throttling  loss  of  the  una-flow  engine  with  single- 
beat  valves  is  smaller  than  that  of  any  other  steam  engine.  In  the  first  place,  the 
throttling  losses  occurring  between  the  cylinders  of  multi-stage  engines  are  eli- 
minated altogether.  The  piston-controlled  exhaust  permits  of  a  sufficient  port 
area  even  with  very  small  exhaust  lead,  so  that  the  pressures  between  engine 
cylinder  and  condenser  may  equalize  fully.  The  throttling  loss  at  the  toe  of  the 
diagram  is  therefore  reduced  to  a  minimum,  and  the  indirect  loss  due  to  thrott- 
ling, consisting  of  a  loss  of  diagram  area  along  the  compression  line,  is  eliminated. 
The  early  cut-offs  employed  result  in  small  throttling  losses  in  any  case  and  these 
are  still  further  reduced  by  the  use  of  single-beat  valves  which  produce  a  com- 
pact stream  instead  of  the  split-up  stream  of  the  ordinary  double-beat  valve,  and 
furthermore  permit  a  fairly  correct  nozzle  diffusor  action  to  be  obtained. 

It  was  also  shown  that  the  leakage  losses  of  the  una-flow  engine  with  single- 
beat  valves  must  be  very  small.  In  the  first  place  the  number  of  points  of  pos- 
sible leakage  is  reduced  to  a  minimum.  While  on  the  one  hand  the  ordinary 
tandem  counterflow  engine  has  three  piston  rod  packings,  eight  valve  stems,  two 
piston  seals  and  sixteen  valve  seats,  in  the  una-flow  engine  with  single-beat  valves 
these  are  reduced  to  one  piston  rod  packing,  two  valve  stems,'  one  piston  seal 
and  two  valve  seats.  Piston  rod  stuffing  boxes  can  be  made  perfectly  tight  by 
use  of  metallic  packings  of  modern  design.  Similarly,  leakage  past  the  valve 
stems  can  be  completely  prevented  if  they  are  properly  fitted.  A  self-supporting 
piston  can  be  made  perfectly  tight  if  properly  designed,  i.  e.  if  the  outer  rings 
do  not  overrun  the  cylinder  bore,  if  a  sufficient  number  of  rings  is  employed,  if 
they  are  secured  against  creeping,  and  their  joints  are  placed  at  the  lowest  point 
so  that  the  part  of  the  piston  in  contact  with  the  cylinder  wall  prevents  the  steam 
from  reaching  the  joints.  From  Fig.  6,  Chapter  VI,  it  will  also  be  seen  that  during 
the  periods  of  high  pressures  all  six  rings  are  active  in  forming  a  seal,  and  later 
on  when  the  pressure  has  fallen  appreciably  three  rings  still  remain  active.  Finally 
the  small  high-lift  single-beat  valve  (see  Fig.  4,  Chapter  VI)  will  remain  perfectly 
tight,  since  its  diameter  is  small,  there  is  only  one  seat,  and  the  sealing  pressure 
is  proportional  to  the  pressure  difference.  Here  the  superiority  in  regard  to 
tightness  of  the  una-flow  engine  with  single-beat  valves  will  find  its  strongest 
expression.  Last  but  not  least,  the  series  arrangement  of  live  steam  space, 
inlet  valve,  piston  and  exhaust  is  also  a  very  valuable  feature  of  this  type  of 
engine.  It  should  therefore  be  possible  to  attain  complete  tightness  at  all  points. 

It  was  shown  that  the  una-flow  engine  with  single-beat  valves  has  the  smallest 
radiation  and  convection  losses. 

Since  the  una-flow  engine  with  single-beat  valves  possesses  the  smallest 
losses  due  to  surface,  volume,  friction,  throttling,  leakage  and  radiation,  there- 
fore very  small  mean  effective  pressures  are  theoretically  permissible. 


318 

It  is  perfectly  clear  that  the  favorable  thermal  action  of  the  surfaces  of  the 
una-flow  engine  with  single-beat  valves  makes  small  mean  effective  pressures 
economically  possible. 

Exchanging  in  rule  7  on  page  42  the  words  "back  pressure"  and  "mean 
effective  pressure",  then  the  rule  reads:  "For  a  given  amount  of  initial  pressure^ 
back  pressure,  clearance  volume  and  length  of  compression,  the  mean  effective  pres- 
sure must  be  chosen  in  such  a  way  as  to  make  the  change  of  total  heat  during  com- 
pression equal  to  the  change  of  total  heat  during  expansion."  If  now  the  clearance 
volume  of  the  una-flow  engine  with  single-beat  valves  is  about  1%,  then  the 
terminal  compression  pressure  will  very  closely  approach  the  initial  pressure,  and 
in  accordance  with  the  above  rule  this  requires  expansion  to  the  back  pressure 
and  mean  effective  pressure  equal  to  zero. 

No  other  type  of  engine  gives  such  fine,  sharp-cornered,  no-load  cards  free 
from  throttling  as  the  una-flow  engine  with  single-beat  valves.  This  type  of 
engine  is  therefore  especially  advantageous  in  cases  where  long  periods  of  idle 
running  are  unavoidable,  as  for  instance  in  rolling  mill  engines.  The  small  thrott- 
ling losses  therefore  also  favor  low  mean  effective  pressures. 

It  is  clear  without  further  comment  that  small  friction  losses  make  small 
mean  effective  pressures  permissible. 

Valve  leakage  in  una-flow  engines  produces  a  certain  increase  in  terminal  pres*- 
sure  at  the  end  of  expansion  and  compression  and  a  corresponding  loss  in  area 
of  the  diagram.  Piston  leakage  on  the  other  hand  results  in  a  loss  of  pressure  at 
the  end  of  expansion  and  compression,  also  with  an  equivalent  loss  of  area.  Both 
kinds  of  losses  increase  with  increasing  ratio  of  expansion  or  compression,  or 
decreasing  mean  effective  pressure.  The  perfectly  tight  steam  distributing  ele- 
ments of  the  una-flow  engine  with  single-beat  valves  therefore  make  small  mean 
effective  pressures  feasible. 

It  is  also  obvious  that  small  losses  due  to  radiation  and  convection  permit 
the  use  of  low  mean  effective  pressures. 

Hence  the  reduction  to  a  minimum  of  all  the  six  losses  so  far  discussed  makes 
it  possible  to  work  with  low  mean  effective  pressures.  Since  a  low  mean  effective 
pressure  is  accompanied  by  a  small  loss  due  to  incomplete  expansion,  this  leads 
to  the  following  statement:  The  reduction  to  a  minimum  of  the  first  six  losses  has 
as  its  consequence  a  minimum  of  the  seventh  loss,  i.  e.  a  minimum  loss  due  to  incom- 
plete expansion. 

Small  mean  effective  pressures,  however,  result  in  larger  cylinder  dimensions, 
therefore  higher  piston  loads  and  higher  first  cost,  the  latter  to  a  greater  extent 
than  in  multi-stage  engines.  This  finally  leads  to  a  compromise  between  the  re- 
quirements for  high  economy  and  low  initial  cost,  i.  e.  the  use  of  higher  mean 
effective  pressures  in  practice,  with  a  somewhat  larger  loss  due  to  incomplete  ex- 
pansion, which  is  on  the  whole  greater  than  that  of  multi-stage  engines.  In  regard 
to  the  loss  due  to  incomplete  expansion,  the  una-flow  engine  with  single-beat  valves 
therefore  ranks  high  theoretically,  but  practical  considerations  forbid  the  full  realization 
of  this  advantage. 

In  the  chapter  on  the  loss  due  to  incomplete  expansion  ways  and  means  were 
indicated  by  which  this  loss  may  be  considerably  reduced  with  a  simultaneous 


319 

increase  in  the  mean  effective  pressure.  This  has  led  to  the  successful  develop- 
ment of  the  una-flow  engine  with  exhaust  ejector  action  as  typified  by  the  una- 
flow  locomotive  with  single-beat  valves,  in  which  the  exhaust  energy  of  one  cylinder 
is  used  to  create  a  vacuum  in  the  second  cylinder.  This  proves  that  in  the  case 
of  multi-cylinder  una-flow  engines  the  loss  due  to  incomplete  expansion  may  be 
minimized,  and  this  leaves  hope  that  the  same  result  may  also  be  accomplished 
for  the  other  kinds  of  service  for  which  the  una-flow  engine  is  so  well  adapted. 

Finally  it  may  be  claimed  that  in  the  una-flow  engine  with  single-beat  valves 
and  double-speed  lay  shaft,  all  the  losses  are  a  minimum  except  that  due  to  incom- 
plete expansion,  with  the  possibility  that  in  the  future  this  loss  also  may  be  reduced 
to  a  minimum. 

Since  the  first  cost  of  a  una-flow  engine  is  usually  15%  lower  than  that  of 
the  equivalent  tandem  compound  engine,  it  would  be  correct  to  reduce  the  mean 
effective  pressure  of  the  former  by  such  an  amount  that  this  difference  in  first 
cost  is  wiped  out.  In  most  cases,  however,  the  striving  after  a  reduction  in  first 
cost  restrains  the  designer  from  availing  himself  of  this  possibility. 

The  uni-directional  flow,  single  stage  expansion,  piston-controlled  exhaust 
and  single-beat  inlet  valves  are  common  features  of  both  the  new  una-flow 
steam  engine  and  the  two  stroke  internal  combustion  engine,  while  the  uni- 
directional flow  is  also  a  feature  of  the  steam  turbine.  Thus  a  certain  similarity 
is  established  in  the  design  and  performance  of  the  two  stroke  internal  com- 
bustion engine  and  the  new  una-flow  steam  engine,  thereby  proving  conclusively 
that  the  latter  rests  upon  sound  principles. 


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